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TJ 1061 
,B4 
1912 
Copy 1 

■ § BEARINGS 

DESIGN—FRICTION-LUBRICATION— 
BEARING METALS 


THIRD EDITION-REVISED AND ENLARGED 












MACHINERY’S REFERENCE BOOKS 


This treatise is one unit in a comprehensive Series of Reference hooks originated 
by Machinery, and including an indefinite number of compact units, each covering 
one subject thoroughly. The whole series comprises a complete working library 
of mechanical literature. The price of each books is 25 cents (one shilling) de¬ 
livered anywhere in the world. 


LIST OF REFERENCE BOOKS 


No. 1. Worm Gearing. —Calculating Di¬ 
mensions; Hobs; Location of Pitch Cir¬ 
cle; Self-Locking Worm Gearing, etc. 

No. 2. Drafting-Room Practice .— 
Systems; Tracing, Lettering and Mount¬ 
ing. 

No. 3. Drill Jigs. —Principles of Drill 
Jigs; Jig Plates; Examples of Jigs. 

No. 4. Milling Fixtures.-— Principles of 
Fixtures; Examples of Design. 

No. 5. First Principles of Theoretical 
Mechanics. 

No. 6. Punch and Die Work.— Princi¬ 
ples of Punch and Die Work; Making and 
Using Dies; Die and Punch Design. 

No. 7. Lathe and Planer Tools. —Cut¬ 
ting Tools; Boring Tools; Shape of Stan¬ 
dard Shop Tools; Forming Tools. 

No. 8. Working Drawings and Draft¬ 
ing Room Kinks. 

No. 9. Designing and Cutting Cams.— 

Drafting of Cams; Cam Curves; Cam De¬ 
sign and Cam Cutting. 

No. 10. Examples of Machine Shop 
Practice. —Cutting Bevel Gears; Making 
a Worm-Gear; Spindle Construction. 

No. 11. Bearings. —Design of Bear¬ 
ings; Causes of Hot Bearings; Alloys 
for Bearings; Friction and Lubrication. 

No. 12. Out of print. 

No. 13. Blanking Dies. —Making Blank¬ 
ing Dies; Blanking and Piercing Dies; 
Split Dies; Novel Ideas in Die Making. 

No. 14. Details of Machine Tool De¬ 
sign. —Cone Pulleys and Belts; Strength 
of Countershafts; Tumbler Gear Design; 
Faults of Iron Castings. 

No. 15. Spur bearing. —Dimensions; 
Design; Strength; Durability. 

No. 16. Machine Tool Drives. —Speeds 
and Feeds; Single Pulley Drives; Drives 
for High Speed Cutting Tools. 

No. 17. Strength of Cylinders. —For¬ 
mulas, Charts, and Diagrams. 

No. 18. Shop Arithmetic for the Ma¬ 
chinist. —Tapers; Change Gears; Cutting 
Speeds; Feeds; Indexing; Gearing for Cut¬ 
ting Spirals; Angles. 

No. 19. Use of Formulas in Mechanics. 
—With numerous applications. ' * • 

No. 20. Spiral Gearing. —Rules, J^tfrtnu- 
las, and Diagrams, etc. 

No. 21. Measuring Tools. —History of 
Standard Measurements; Calipers; Com¬ 
passes; Micrometer Tools; Protractors. 

No. 22. Calculation of Elements of 
Machine Design. —Factor of Safety; 
Strength of Bolts; Riveted Joints; Keys 
and Key ways; Toggle-joints. 

No. 23. Theory of Crane Design.— Jib 
Cranes; Shafts, Gears, and Bearings; 
Force to Move Crane Trolleys; Pillar 
Cranes. 


No. 24. Examples of Calculating De¬ 
signs. —Charts in Designing; Punch and 
Riveter Frames; Shear Frames; Billet 
and Bar Passes; etc. 

No. 25. Deep Hole Drilling. —Methods 
of Drilling; Construction of Drills. 

No. 26. Modern Punch and Die Con¬ 
struction. —Construction and Use of Sub¬ 
press Dies; Modern Blanking Die Con¬ 
struction; Drawing and Forming Dies. 

No. 27. Locomotive Design, Part I. — 
Boilers, Cylinders, Pipes and Pistons. 

No. 28. Locomotive Design, Part IT.— 
Stephenson and Walschaerts Valve Mo¬ 
tions; Theory, Calculation and Design. 

No. 29. Locomotive Design, Part III. 
—Smokebox; Exhaust Pipe; Frames; 
Cross-heads; Guide Bars; Connecting-rods; 
Crank-pins; Axles; Driving-wheels. 

No. 30. Locomotive Design, Part IV.— 
Springs, Trucks, Cab and Tender. 

No. 31. Screw Thread Tools and Gages. 

No. 32. Screw Thread Cutting. —Lathe 
Change Gears; Thread Tools; Kinks. 

No. 33. Systems and Practice of the 
Drafting-Room. 

No. 34. Care and Repair of Dynamos 
and Motors. 

No. 35. Tables and Formulas for Shop 
and Drafting-Room. —The Use of Formu¬ 
las; Solution of Triangles; Strength of 
Materials; Gearing; Screw Threads; Tap 
Drills; Drill Sizes; Tapers; Keys, etc. 

No. 36. Iron and Steel. —Principles of 
Manufacture and Treatment. 

No. 37. Bevel Gearing. —Rules and 
Formulas; Examples of Calculation; 
Tooth Outlines; Strength and Durability; 
Design; Methods of Cutting Teeth. 

No. 38. Out of print. See No. 98. 

No. 39. Fans, Ventilation and Heating. 
—Fans; Heaters; Shop Heating. 

No. 40. Fly-Wheels. —T heir Purpose, 
Calculation and Design. 

No. 41. Jigs and Fixtures, Part I. — 
Principles of Design; Drill Jig Bushings; 
Locating Points; Clamping Devices. 

No. 42. Jigs and Fixtures, Part II.— 
Open and Closed Drill Jigs. 

No. 43. Jigs and Fixtures, Part III. — 
Boring and Milling Fixtures. 

No. 44. Machine Blacksmithing. —Sys¬ 
tems, Tools and Machines used. 

No. 45. Drop Forging. — Lay-out of 
Plant; Methods of Drop Forging; Dies. 

No. 46. Hardening and Tempering.— 
Hardening Plants; Treating High-Speed 
Steel; Hardening Gages. 

No. 47. Electric Overhead Cranes.— 
Design and Calculation. 

No. 48. Files and Filing. —Types of 
Files; Using and Making Files. 

No. 49. Girders for Electric Overhead 
Cranes. 


(See inside back cover for additional titles) 


MACHINERY’S REFERENCE SERIES 


EACH NUMBER IS ONE UNIT IN A COMPLETE LIBRARY OF 
MACHINE DESIGN AND SHOP PRACTICE REVISED AND 
REPUBUSHED FROM MACHINERY 


NUMBER 11 

BEARINGS 

DESIGN—FRICTION—LUBRICATION—BEARING METALS 
Third Edition—Revised and Enlarged 

CONTENTS 

The Design of Bearings, by Forrest E. Cardullo - 3 

Causes of Hot Bearings, by E. Kistinger 15 

Thrust Bearings.21 

Friction and Lubrication.25 

Bearing Metals, by Joseph H. Hart - 35 

Alloys for Bearings.40 

Friction of Roller Bearings, by C. H. Benjamin - - 45 

PROPERTY-OF 

library of congress 


Copyright, 1912, The Industrial Press, Publishers of Machinery, 
49-55 Lafayette Street, New York City 




In the second edition of this Reference Series Book, a chapter on 
the principles of thrust bearings was introduced, together with a very 
important chapter on friction and lubrication of bearings. An addi¬ 
tional chapter on bearing metals was also included in that edition. In 
order to provide space for this material, the chapter on ball bearings 
which was contained in the first edition, was eliminated. This chapter, 
together with a considerable amount of additional matter on the 
same subject, is included in Machinery’s Reference Book, No. 56, “Ball 
Bearings.” 




©Cl. A327953 
T-t-© • /, 


CHAPTER I 


THE DESIGN OF BEARINGS* 

The design of journals, pins, and bearings of all kinds is one of the 
most important problems connected with machine construction. It is 
a subject upon which we have a large amount of data, but, unfor¬ 
tunately, they are very conflicting. The results obtained from the rules 
given by different mechanical writers will be found to differ by GO 
per cent or more. Many of our best modern engines have been designed 
in defiance of the generally accepted rules on this subject, and many 
other engines, when provided with what were thought to be very liberal 
bearing surfaces have proved unsatisfactory. This confusion has 
largely been the result of a misconception of the actual running con¬ 
ditions of a bearing. 

Friction of Journals 

A journal should be designed of such a size and form that it will run 
cool, and with practically no wear. The question both of heating and 
wear is one of friction, and in order for us to understand the princi¬ 
ples upon which the design of bearings should be based, we must first 
understand the underlying principles of friction. Friction is defined 
as that forcef acting between two bodies at their surface of contact, 
when they are pressed together, which tends to prevent their sliding 
one upon the other. The energy used in overcoming this force of 
friction, appears at the rubbing surfaces as heat, and is ordinarily 
dissipated by conduction through the two bodies. The force of friction, 
and hence the amount of heat generated under any given circumstances, 
can be greatly reduced by the introduction of an oily or greasy sub¬ 
stance between the rubbing surfaces. The oil or grease seems to act 
in the same way that a great number of minute balls would, reducing 
the friction and wear, and thus preventing the overheating and conse¬ 
quent destruction of the parts. On this account, bearings of all kinds 
are always lubricated. Thus the question of journal friction involves 
the further question of lubrication. 

For the purpose of understanding as far as possible what goes on 
in a bearing, and the amount and nature of the forces acting under 
different conditions, several machines have been designed to investi¬ 
gate the matter. In general they are so arranged that a journal may 
be rotated at any desired speed, with a known load upon the boxes. 
Suitable means are provided for measuring the force of friction, and 
also the temperature of the bearing. .Provided with such an apparatus, 
we find that the laws of friction of lubricated journals differ very 
materially from those commonly stated in the text-books as the laws 
of friction. A comparison of the two will prove interesting. 

* Machinery, December, 1906 ; January and February, 1907. 

t Friction. * * * Resistance to motion due to the contact of surfaces.— 

Standard Dictionary. —Force. * * * Any cause that produces, stops, changes, 

or tends to produce, stop, or change the motion of a body .—Standard Dictionary . 



4 


No. ii—BEARINGS 

Frictional Resistance in Lubricated and Unlubricated Bearing-s 

It is generally stated in the text-hooks that the force of friction is 
proportional to the force with which the rubbing surfaces are pressed 
together, doubling, or trebling, as the case may be, with the normal 
pressure. This law is perfectly true for all cases of unlubricated 
bearings, or for bearings lubricated w r ith solid substances, such as 
graphite, soapstone, tallow, etc. When, however, the bearing is prop¬ 
erly lubricated with any fluid, it is found that doubling the pressure 
does not by any means double the friction, and when the lubricant 
is supplied in large quantities by means of an oil bath or a force 
pump, the friction will scarcely increase at all, even when the pressure 
is greatly increased. From the experiments of Prof. Thurston, and 
also of Mr. Tower, it appears that the friction of a journal per square 
inch of bearing surface, for any given speed, is equal to 

/ = kp n (1) 

where / is the force of friction acting on every square inch of bearing 
surface, p is the normal pressure in pounds per square inch on that 
surface, and k is a constant. The exponent n depends on the manner 
of oiling, and varies from 1 in the case of dry surfaces, to 0.50 in the 
case of drop-feed lubrication, 0.40 or thereabouts in the case of ring- 
and chain-oilers and pad lubrication, and becomes zero in case the oil 
is forced into the bearing under sufficient pressure to float the shaft. 

The second law of friction, as generally stated, is that the force of 
friction is independent of the velocity of rubbing. This law also is 
true for unlubricated surfaces, and for surfaces lubricated by solids. 
In the case of bearings lubricated by oil we find that the friction 
increases with the speed of rubbing, but, not at the same rate. If we 
express the law as an equation, we have 

f = kv m (2) 

where / is the force of friction at the rubbing surfaces in pounds per 
square inch, k is a constant, v is the velocity of rubbing in feet per 
second, and the exponent m varies from zero in the case of dry surfaces 
to 0.20 in the case of drop-feed, and Q.50 in the case of an oil bath. 

The third law of friction, as it generally appears in the text-books, 
is that the friction depends, among other things, on the composition of 
the surfaces rubbed together. This, again, while true for unlubricated 
surfaces, is not true for other conditions. It matters nothing whether 
the surfaces be steel, brass, babbitt, or cast iron, so long as they are 
perfectly smooth and true, they will have the same friction when 
thoroughly lubricated. The friction will depend upon the oil used, not 
on the materials of journal or boxes, when the other conditions of 
speed and pressure remain constant. Many people think that babbitt 
has less friction than iron or/brass, under the same circumstances, 
but this is not true. The reason for the great success of babbitt as an 
“anti-friction” metal depends upon an entirely different property, as 
will appear later. 

Combining into one equation the different laws of the friction of 
lubricated surfaces, as we actually find them to be, we have 

f — kp*v m 


( 3 ) 


DESIGN OF BEARINGS 


5 


where / is the force of friction at the rubbing surface in pounds per 
square inch, Jc is a constant which varies with the excellence of the 
lubricant from 0.02 to 0.04, and the other quantities are as before. 
From this expression, we see that the friction increases with the load 
on the bearing, and also with the velocity of rubbing, although much 
more slowly than either. 

Generation of Heat in Bearings 

The quantity of heat generated per square inch of bearing area, 
per second, is equal to the force of friction, times the velocity of rub¬ 
bing. All of this heat must be conducted away through the boxes as 
fast as it is generated, in order that the bearing shall not attain a 
temperature high enough to destroy the lubricating qualities of the oil. 
The hotter the boxes become, the more heat they will radiate in a 
given time. When the bearing is running under ordinary working 
conditions, it will warm up until the heat radiated equals the heat 
generated, and the temperature so attained will remain constant as 
long as the conditions of lubrication, load, and speed do not change. 
This rise in temperature above that of the surrounding air varies from 
less than 10 to nearly 100 degrees Fahrenheit, and is commonly about 
SO degrees. We must keep either the force of friction or the velocity 
of rubbing, or both, down to that point where the temperature shall not 
attain dangerous values. As has been shown in the preceding para¬ 
graph, it was formerly believed that the force of friction was equal 
to a constant times the bearing pressure, and therefore, that the work 
of friction was equal to this constant times the pressure, times the 
velocity of rubbing. Now, since it is the work of friction that we are 
obliged to limit to a certain definite value per square inch of bearing 
area, it was concluded that a bearing would not reach a dangerous 
temperature if the product of the bearing pressure per square inch and 
the velocity of rubbing did not exceed a certain value. Accordingly, 
we find Prof. Thurston’s formula for bearings to be 

pv~C. (4) 

where p is the bearing pressure in pounds per square inch, v is the 
velocity of rubbing in feet per second, and C has values varying from 
800 foot-pounds per second in the case of iron shafts to 2,600 in the 
case of steel crank-pins. This has long been the standard formula for 
designing bearings, and while it is not satisfactory in extreme cases, 
it is very satisfactory for bearings running at ordinary speeds. 

Turning our attention again to the results obtained from the ma¬ 
chines for testing bearings, we find that while the results are very 
even and regular for ordinary pressures and temperatures, when we 
begin to increase either of these to a high point, the friction and wear 
of our bearing suddenly increases enormously. The reason is that the 
oil has been squeezed out of the bearing by the great pressure. This 
squeezing out of the oil, and consequent great increase in the friction, 
has three effects. The absence of the lubricant causes the parts to 
scratch or score each other, thus rapidly destroying themselves, the 
great increase in friction results in a sudden very high temperature, 


6 


No. ii—BEARINGS 


in itself destructive to the materials of the bearing, and the heating 
is generally so rapid as to cause the pin and the interior parts of the 
box to expand more rapidly than the exterior parts, thus causing the 
box to grip the pin with enormous pressure. When the oil has been 
squeezed out in this manner, the bearing is said to seize. 

Materials for Bearing-s 

It is evidently of advantage to make the bearing of such material 
that the injury resulting from seizing shall be a minimum. If the 
shaft and box are of nearly equal hardness, each will tend to scratch 
the other when seizing occurs, and the scoring is rapid and destructive. 
This action will be especially noticed in case the shaft has hard spots 
in it, while the rest is comparatively soft, as is the case in the poorer 
grades of wrought iron. If, however, the shaft is made of a hard 
and homogeneous material, like the better grades of medium steel, and 
the bearing is made of some soft material, like babbitt, the bearing 
will not roughen the journal, and so the journal cannot cut the bearing. 
This is the first reason why babbitt bearings are so successful. 

A second reason for the success of babbitt bearings lies in the fact 
that they cannot be heated sufficiently to make the bearing grip the 
journal. They will rather soften and flow under the pressure without 
actually melting away, just as iron and steel soften at a welding heat. 
The harder bearing metals, such as brass and bronze, do not have these 
advantages, and have been almost entirely replaced by babbitt in bear¬ 
ings for heavy duty, especially when thorough lubrication is difficult. 

Babbitt is a successful bearing metal for still a third reason. The 
unit pressure on any bearing is not the same at all points. The shaft 
is invariably made somewhat smaller in diameter tha» the box. If 
there is a high spot on the surface of the box, that spot will have a 
very large proportion of the total pressure acting on it, and as a result 
the film of lubricant will be broken down at that point, and local 
heating and consequent damage result. In the case of babbitt bearings, 
before the damage can become serious the metal is caused to flow 
away from that point under the combined influence of the heat and 
pressure, the oil film is again established, and normal conditions 
restored. 

Influence of Quality of Oil 

The unit pressure which any bearing will stand without seizing 
depends upon its temperature and the kind of oils used. The lower 
the temperature of the bearings, the greater the allowable unit pres¬ 
sure. The reason for this is that oils become thinner and more free- 
flowing at the higher temperatures, consequently they are more easily 
squeezed out of the bearing, and it is more likely to seize. On this 
account, the higher the velocity of rubbing, the less the unit pressure 
that can be carried, but it does not follow that the allowable unit 
pressure varies inversely as the speed of rubbing, as was formerly 
thought. 

The thicker and less free-flowing an oil is, the greater the unit 
pressure it will stand in a bearing without squeezing out. A watch 


DESIGN OF BEARINGS 


7 


oil, or a very light spindle oil, will only run under a very small unit 
pressure; sometimes they are squeezed out of the hearing when the 
pressure does not exceed 50 pounds per square inch. On the other hand, 
a cylinder oil of good body will stand a pressure of over 2000 pounds 
to the square inch in the same bearing. There is a certain quality of 
oil which is best adapted to every bearing, and if possible it should be 
the one used. 

A third cause influencing the pressure which may be oarried is 
adhesiveness between the oil and the rubbing surfaces. Some oils are 
more certain to wet metal surfaces than are others, and in the same 
way some metals are more readily wet by oil than are others. It is 
evident that when the surfaces repel, rather than attract, the oil, the 
film will be readily broken down, and when the opposite is the case 
the film is easily preserved. 


Oil Grooving’ 

The mechanical arrangement of the box and journal may tend either 
to preserve or destroy the lubricating film. Both should be perfectly 
round and smooth, the box a trifle larger in diameter than the journal. 
The allowance commonly made for the “running fit” of the box and 
shaft is about 0.0005 (D + 1) inches, where D is the nominal diameter 
of the shaft in inches. Some manufacturers of fast-running machinery 
make the diameter of the box exceed that of the shaft by nearly twice 
this amount. The oil should be introduced at that point where the 
forces acting tend to separate the shaft and box. At this point grooves 
must be cut in the surface of the box, so as to distribute the lubricant 
evenly over the entire length of the journal. Having been so intro¬ 
duced and distributed, the oil will adhere to the journal, and be carried 
around by it as it revolves to the point where it is pressed against the 
box with the greatest force, thus forming the lubricating film which 
separates the rubbing surfaces. The supply of lubricant thus con¬ 
tinually furnished, and swept up to the spot where it is needed, must 
not be diverted from its course in any way. A sharp edge at the 
division point of the box will wipe it off the journal as fast as it is 
distributed, or a wrongly placed oil groove will drain it out before it 
has entirely accomplished its purpose. 

An important matter in the design of bearings is the cutting of these 
oil grooves. They are a necessary evil, and should be treated as such, 
by using as few of them as possible. They serve, first, to distribute 
the lubricant uniformly over the surface of the journal, and, second, 
to collect the oil, which would otherwise run out at the ends of the 
bearing, and return it to some point where it may again be of use. As 
generally cut, oil grooves have two faults; first, they are so numerous 
as to cut down to a serious extent the area of the bearing, and, second, 
they are so located as to allow the oil to drain out of the bearing. 
Let us take an ordinary two-part cap bearing such as the outboard 
bearing of a Corliss engine, and see how it is best to cut the grooves. 

One of these bearings, as commonly made by good builders, is shown 
in Fig. 1. The oil is supplied, drop by drop, through a hole in the 


8 


No. ii—BEARINGS 

cap. If there were no oil grooves, only a narrow band of the shaft 
revolving immediately under this hole would he reached by the oil. 
If now, we cut a shallow groove in the cap, lengthwise of the bearing, 
and reaching almost, but not quite, to the edges, the oil will be enabled 
to reach every part of the revolving surface. To this groove we some¬ 
times add two, as shown by the dotted lines in Fig. 2, which show the 
inner surface of the cap as being unrolled, and lying flat on the paper. 
No series of grooves can be cut in the box which will distribute the 
oil as well or as thoroughly as those shown, and they should always 
be used in the caps of such bearings in preference to any others. 

Having distributed the oil over the revolving surface, our next care 
must be to see that it is not wiped off before it reaches the point for 
which it was intended. Accordingly, we should counterbore the box 
at the joint in such a way as to make a recess in which the surplus oil 
may gather, and which will further assist when necessary in dis¬ 
tributing the lubricant. This counterbore should extend to within r /\ 
or y 2 inch of the ends of the bearing, as shown in Fig. 1. 

When the oil is supplied through the cap, grooves for the distribution 




Fig. 1. Section of Outboard Bearing, showing Oil Grooving and 
Counterbored Recess 

of the oil should not be cut in the bottom half of the bearing, since 
they will only serve to drain the bearing of the film of oil formed 
there. The old film is under great pressure at this point, and naturally 
tends to flow away when any opportunity is offered. If left to its own 
devices, part of it will squeeze out at the ends of the bearing and be 
lost. In order to save this oil, shallow grooves, parallel to the ends 
of the bearing, may be cut in the lower box, as shown in Figs. 1 and 3. 
Their office is to intercept the oil which would flow out at the ends, 
and divert it to the counterbored recesses, where it can again be made 
of use. These are the only grooves that should ever be used in the 
lower half of a two-part bearing, and they should only be used in the 
larger sizes. 

Two classes of bearings which may well be made without oil grooves 
are, first, the cross-head slippers of engines, and, second, crank-pin 
boxes. The cross-head slipper should have a recess cut at each end, 
in the same way as the counterboring of the two-part box, as shown in 
Fig. 4. To this is sometimes added the semi-circular groove shown in 


























DESIGN OF BEARINGS 


9 


dotted lines, which does no harm, although it is unnecessary. The best 
way to oil a crank-pin is through the pin itself. In the case of over¬ 
hung pins, a hole is drilled lengthwise of the pin to its center. A 
second hole is drilled from the surface of the pin to meet the first one. 
A shallow groove should now be cut in the surface of the pin, parallel 
to its axis, and reaching almost to the ends of the bearing, as shown 



Fig. 2. Development of Cap, showing Oil Grooving and 
Counterboring 

in Fig. 5. No grooves should be cut in the boxes, but the edges where 
they come together should be counterbored. 

As much care and attention should be given to the oil grooving as 
to the size of a bearing, yet it is a matter often left to the fancy of 
the mechanic who fits it. The purpose of the grooves, to distribute the 
oil evenly, should ever be kept in mind, and no groove should be cut 
which does not accomplish, this purpose, except it be to return waste 



J Machinery, N. Y. 


Fig. 3. Development of Lower Half of Outboard Bearing 

oil to a place where it may again be of use. Most commonly, bearings 
have too many grooves. So far from helping the lubricants, they 
generally drain the oil from where it is most needed. Use them 
sparingly. 

Calculating- the Dimensions 

The durability of the lubricating film is affected in great measure 
by the character of the load that the bearing carries. When the load 
is unvarying in amount and direction, as in the case of a shaft carry¬ 
ing a heavy fly-wheel, the film is easily ruptured. In those cases 
where the pressure is variable in amount and direction, as in railway 
journals and crank-pins, the film is much more durable. When the 
journal only rotates through a small arc, as with the wrist-pin of a 




























10 


No. n—BEARINGS 


steam engine, the circumstances are most favorable. It has been 
found that when all other circumstances are exactly similar, a car 
journal, where the force varies continually in amount and direction, 
will stand about twice the unit pressure that a fly-wheel journal will, 
where the load is steady in amount and direction. A crank-pin, since 
the load completely reverses every revolution, will stand three times, 
and a wrisit-pin, where the load only reverses, but does not make a 
complete revolution, will stand four times the unit pressure that the 
fly-wheel journal will. 

The amount of pressure that commercial oils will endure at low 
speeds without breaking down varies from 500 to 1000 pounds per 
square inch, where the load is steady. It is not safe, however, to 
load a bearing to this extent, since it is only under favorable circum¬ 
stances that the film will stand this pressure without rupturing. On 
this account, journal bearings should not be required to stand more 
than two-thirds of this pressure at slow speeds, and the pressure 







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Machinery, N.Y, 

Figr. 4. Face of Cross-head Slipper 


should be reduced when the speed increases. The approximate unit 
pressure which a bearing will endure without seizing is as follows: 

P K 

P = - (5) 

DN + K 

where p is the allowable pressure in pounds per square inch of pro¬ 
jected area, D is the diameter of the bearing in inches, N is the num¬ 
ber of revolutions of the journal per minute, and P and K depend upon 
the kind of oil, manner of lubrication, etc. 

The quantity P is the maximum safe unit pressure for the given 
circumstances, at a very slow speed. In ordinary cases the value of 
this number will be 200 for collar thrust bearings, 400 for shaft bear¬ 
ings, 800 for car journals, 1200 for cank-pins, and 1600 for wrist-pins. 
In exceptional circumstances, these values may be increased by as 
much as 50 per cent, but only when the workmanship is of the best, 
the care the most skillful, the bearing readily accessible, and the oil 
of the best quality and unusually viscous. It is only in the case of 
very large machinery, which will have the most expert supervision, 
that such values can be safely adopted. In the case of the great units 
built for the Subway power plant in New York by the Allis-Chalmers 










DESIGN OF BEARINGS 


11 


Co. the value of P in the formula given on page 10 for the crank-pins 
was 2,000—as high a value as it is ever safe to use. 

The factor K depends upon the method of oiling, the rapidity of 
cooling, and the care which the journal is likely to get. It will be 
found to have about the following values: Ordinary work, drop-feed 
lubrication, 700; first-class care, drop-feed lubrication, 1,000; force- 
feed lubrication or ring-oiling, 1,200 to 1,500; extreme limit for perfect 
lubrication and air-cooled bearings, 2,000. The value 2,000 is seldom 
used, except in locomotive work where the rapid circulation of the 
air cools the journals. Higher values than this may only be used in 
the case of water-cooled bearings. 

Formula No. 5 is in a convenient form for calculating journals. 
In case the bearing is some form of a sliding shoe, the quantity 240 V 
should be substituted for the quantity D N in the equation, V being 
the velocity of rubbing in feet per second. There are few cases 
where a unit pressure sufficient to break down the oil film is ailOW- 




Fig. 6. Section showing the Bend¬ 
ing of a Crank-pin and Consequent 
Unequal Wear of the Box 


able. Such cases are the pins of punching and shearing machines, 
pivots of swing bridges, and so on. The motion is so slow that heat¬ 
ing cannot well result, and the effects of scoring cannot be serious. 
Sometimes bearing pressures up to the safe working stress of the 
material are used, but better practice is to use pressures not in excess 
of 4,000 pounds per square inch. 

In general, the diameter of a shaft or pin is fixed from considera¬ 
tions of strength or stiffness. Having obtained the proper diameter, 
we must next make the bearing long enough so that the unit pressure 
shall not exceed the required value. This length may be found 
directly by means of the equation: 








































12 


No. ii— BEARINGS 


where L is the length of the bearing in inches, W the load upon it in 
pounds, and P, K, N, and D are as before. 

A bearing may give poor satisfaction because it is too long, as well 
as because it is too short. Almost every bearing is in the condition 
of a loaded beam, and therefore it has some deflection. Let us take 
the case of an overhung crank-pin, in order to examine the phenomena 
occurring in a bearing under these circumstances. When the engine 
is first run, both the pin and box are, or should be, truly round and 
cylindrical. As the pin deflects under the action of the load, the 
pressure becomes greater on the side toward the crank throw, break¬ 
ing down the oil film at that point, and causing heat. After a while 
the box becomes worn to a slightly larger diameter at the side toward 
the crank, in the manner shown in Fig. 6, which is a section showing 
an exaggerated view of the condition of affairs in the crank-pin box 
when under load. 

It has already been noted that the box must be a trifle larger in 
diameter than the journal, and for successful working this difference 
is very strictly defined, and can vary only within narrow limits. 
Should the pin be too large, the oil film will be too thin, and easily 
ruptured. On the other hand, should the pin be too small the bearing 
surface becomes concentrated at a line, and the greater unit pressure 
at that point ruptures the film. This is exactly what happens when 
the pin is too long. The box rapidly wears large at the inner end, 
and the pressure becomes concentrated along a line as a consequence. 
The lubricating film then breaks down, and the pin heats and scores. 
The remedy is not to make the pin longer, so as to reduce the unit 
pressure, but to decrease its length and to increase its diameter, caus¬ 
ing the pressure to be evenly distributed over the entire bearing 
surface. 

The same principles apply to the design of shafts and center crank- 
pins. They must not be made so long that they will allow the load 
to concentrate at any point. A very good rule for the length of a 
journal is to make the ratio of the length to the diameter about equal 
to one-eighth of the square root of the number of revolutions per 
minute. This quantity may be diminished by from 10 to 20 per cent 
in the case of crank-pins, and increased in the same proportion in the 
case of shaft bearings, but it is not wise to depart too far from it. 
In the case of an engine making 100 revolutions per minute, the bear¬ 
ings would be by this rule from one and a quarter to one and a half 
diameters in length. In the case of a motor running at 1,000 revolu¬ 
tions per minute, the bearings would be about four diameters long. 
While the above is not a hard and fast rule which must be adhered 
to on all occasions, it will be found to be an excellent guide in all 
cases of doubt. 

The diameter of a shaft or pin must be such that it will be strong 
and stiff enough to do its work properly. In order to design it for 
strength and stiffness it is first necessary to know its length. This 
may be assumed from the following equation: 


DESIGN OF BEARINGS 


13 


20Wr/N 

L = -, (7) 

PK 

where all the quantities are the same as in the preceding equations. 
Having found the approximate length by the use of the above equa¬ 
tion, the diameter of the shaft or pin may be found by any of the 
standard equations given in the different works on machine design. 
It is next in order to recompute the length from formula No. 6, taking 
this new value if it does not differ materially from the one first as¬ 
sumed. If it does, and especially if it is greater than the assumed 
length, take the mean value of the assumed and computed lengths and 
try again. 

Examples of Calculating Dimensions for Bearings 
A few examples will serve to make plain the methods of designing 
bearings by means of these principles. Let us take as the first case 
the collar thrust bearings on a 10-inch propeller shaft, running at 150 
revolutions per minute, and with a thrust of 60,000 pounds. Assuming 
that the thrust rings will be 2 inches wide, their mean diameter will 
be 12 inches. From equation No. 5 we will have for the allowable 
200 X 700 

bearing pressure -, or 56 pounds per square inch. This 

12 x 150 + 700 

will require a bearing of 60,000 -5- 56, or 1070 square inches area. 
Since each ring has an area of 0.7854 (14 2 — 10*), or about 75 square 
inches, the number of rings needed will be 1070 -s- 75, or 14. In case it 
was desirable to keep down the size of this bearing, the constant K 
might have had values as high as 1000 instead of 700. 

Next, we will take the main bearing of a horizontal engine. We 
will assume that the diameter of the shaft is 15 inches, that the weight 
of the shaft, fly-wheel, crank-pin, one half the connecting-rod, and any 
other moving parts that may be supported by the bearings, is 120,000 
pounds, and that two-thirds of this weight comes on the main bearing, 
the remainder coming on the outboard bearing. The engine runs at 
100 revolutions per minute. In this case, W = 80,000 pounds, P = 400 
pounds per square inch, and K depends upon the care and method of 
lubrication. Assuming that the bearing will be flushed with oil by 
some gravity system, and that, since the engine is large, the care will 
be excellent, we will let K — 1500. This gives us for the length of 
the bearing from formula No. 6: 

80,000 / 1500 \ 

L — -1100 H-I = 26^ inches (about). 

400 X 1500 \ 15 / 

It is to be noted that, in computing the length of this bearing, the 
pressure of the steam on the piston does not enter in, since it is not 
a steady pressure, like the weight of the moving parts. The only 
matter to be noted in connection with the steam load is that the pro¬ 
jected area of the main bearing of an engine shall be in excess of 
the projected area of the crank-pin. 

For another example we will take the case of the bearings of a 






14 


No. ii—BEARINGS 


100,000*pound hopper car, weighing 40,000 pounds, and with eight 33- 
inch wheels. The journals are 5% inches diameter, and the car is to 
run at 30 miles per hour. The wheels will make 307 revolutions per 
minute when running at this speed, and the load on each journal will 
be 140,000 -j-8, or 17,500 pounds. Although the journal will be well 
lubricated by means of an oil pad, it will receive but indifferent care, 
so the value of K will be taken as 1,200. The length of the journal 
will then he 

17,500 / 1,200 \ 

L =- I 307 H - 1=9% inches (about). 

800 X 1,200 \ 5.5 / 

As a last example, we will take the case of the crank-pin of an 
engine with a 20-inch steam cylinder, running at 80 revolutions per 
minute, and having a maximum unbalanced steam pressure of 100 
pounds per square inch. The maximum, and not the mean steam press¬ 
ure should he taken in the case of crank- and wrist-pins. The total 
steam load on the piston is 31,400 pounds. P will be taken as 1,200, 
and E as 1,000. We will therefore obtain for our trial length: 

20 X 31,400 Xj/80 

L = -= 4.7, or, say, 4% inches. 

1,200 X 1,000 

In order that the deflection of the pin shall not be sufficient to de¬ 
stroy the lubricating film we have 

D = 0.09 [/ W L* 

which limits the deflection to 0.003 inch. Substituting in this equa¬ 
tion, we have for the diameter 3.85, or say 3% inches. With this diam¬ 
eter we will obtain the length of the bearing, by using formula No. 
6, and find 

31,400 / 1,000 \ 

L =- I 80 H -I = 8.85, say 9 inches. 

1,200 X 1,000 \ 3% / 

The mean of this value, and the one obtained before is about 7 
inches. Substituting this in the equation for the diameter, we get 
5% inches. Substituting this new diameter in equation No. 6 we have 

31,400 / 1,000 \ 

L =-1 80 H-I = 7.1, say 7 inches. 

1,200 X 1,000 \ 5% / 

Probably most good designers would prefer to take about half an 
inch off the length of this pin, and add it to the diameter, making 
it 5% X 6% inches, and this will be found to bring the ratio of the 
length to the diameter nearer to one-eighth of the square root of the 
number of revolutions. 









CHAPTER II 


CAUSES OP HOT BEARINGS* 

In our modern high-speed steam and gas engines, turbines and tho 
like, hot bearings are of more frequent occurrence than is generally 
supposed. Very often a new plant, just put into service, has to be shut 
down on this account. It not infrequently happens that the engine 
which has run “hot” is one of several, identical in design and con¬ 
struction, the bearings in the others having operated without trouble 
Apparently there is no cause for this particular engine to give trouble, 
but in order to remove the difficulty, various makes of babbitt metals 
and bronzes are tried, sometimes with good results, sometimes without. 
Again, it occurs that a machine or engine operates at the beginning 
with perfect satisfaction, but after a time one or more of the bearings 



Figr, 7. One-piece Bearing” Babbitt just Poured. Both Shell and Babbitt 
at the Solidifying: Temperature 

begin to run “warm,” and finally “hot,” so that relining becomes neces¬ 
sary. As a general rule it is then simply accepted as a fact that the 
bearings “ran hot”; seldom does anyone think it worth while to seek 
out the fundamental causes for the trouble. That there is always the 
element of doubt in regard to bearings, is evidenced by the fact that 
our modern engine builders usually deliver an extra set of bearings 
with the engine, so that, in the event of trouble, a new set is at hand. 
The following may be of some assistance towards discovering and 


"Machinery, November, 1907. 




















16 


No. ii—BEARINGS 

eliminating, in a scientific manner, and along technical and metal¬ 
lurgical lines, the real causes of hot bearings. 

Investigation will show that the main reasons for hot bearings are: 

1. —Shrinkage or contraction of the babbitt. 

2. —Shrinkage strains set up in the babbitt metal liner by the un¬ 
equal distribution of the babbitt metal over the shell. 

3 . —a lack of contact between the babbitt metal liner and the cast 
iron or cast steel shell. 

4 . —The lubricant becomes partially deflected into the wrong place. 

Shrinkage or Contraction of the Babbitt 
a. Shrinkage in a diametral direction. As an illustration of tkis 
point, one may take the simple example of an iron ball and ring. If 
this ball, when cold, will just pass through an iron ring, it will not 



Fig. 8. Same Bearing as shown in Fig. 7 Cooled Down to Normal 
Temperature 

do so when somewhat heated and expanded. After cooling down, how¬ 
ever, it will again pass through the ring. A similar action takes place 
in a bearing. 

In Fig. 7 of the accompanying illustrations the babbitt liner may 
be considered to have been just poured in, and the metal to be still 
liquid. At the exact solidifying point the babbitt will have filled all 
the interstices and be in good contact with the cast iron or cast steel 
shell, provided the babbitt itself has sufficient fluidity to enable it to 
penerate the smallest spaces. From this solidifying point on, the 
babbitt will contract according to its coefficient of contraction. Now, 
if the coefficient of contraction of the babbitt were the same as that of 
the material out of which the shell is made (usually cast iron or cast 

















CAUSES OF HOT BEARINGS 


17 


steel), and provided that the shell had acquired the same temperature 
as the babbitt, the shell and the babbitt liner would then contract 
equally, and a fairly good contact would result, and there would be 
nothing to set up counter strains during shrinkage. But, as the co¬ 
efficient of contraction of almost all babbitt metals is approximately 
two or three times higher than that of cast iron or cast steel, a shrink- 



Machinery, N. Y. 

Fig. 9. Babbitt Bearing without Dove-tailed Grooves or other 
Retaining Device 

age or loosening of the babbitt liner from the shell must absolutely 
take place after the solidfying point of the babbitt is reached. Fig. 8 
shows this contraction as it would appear if magnified. The fact that 
most bearings are “split” does not, of course, change this result. If 
the babbitt is secured in the shell by means of dove-tailed grooves, or 
other anchoring devices, so that the actual visible contraction from the 



Fig. lO. Same Bearing as shown in Fig. 9, but with Dove-taiJed Grooves. 

Visible Shrinkage Prevented, but Shrinkage Strains Produced 

shell is lessened or minimized, then an unavoidable consequence of 
these grooves or other devices is shrinkage strains, set up while the 
babbitt cools down, as explained further on. 

&. Shrinkage in an axial direction. With regard to shrinkage in the 
axial direction, it may be observed that the same results take place. 
Fig. 9 illustrates how the babbitt metal shrinks in a cast iron or cast 
steel shell in the axial direction, when there is no anchoring device 































18 


No. ii—BEARINGS 


employed. In Fig. 10 may be seen the old-fashioned dove-tailed groove 
construction, prohibiting an actual visible shrinkage, but causing 
shrinkage strains. 

Shrinkage Strains Produced by an Unequal Distribution of 
Babbitt Metal Liner 

By referring to Fig. 11, it will be observed that the babbitt metal at 
oa is about twice as thick as at bb. The consequence is that, as the 
solidifying time of the greater mass aa is longer than that of the 
smaller mass bb, shrinkage strains are set up throughout the babbitt 
liner, which loosen it from the shell and have the tendency, in com¬ 
bination with the regular working pressures and shocks, to produce 
minute cracks in the liner. 

Lack of Contact between Liner and Shell 

In a bearing shell some parts of the liner are in close contact with 
the shell, as a result of careful pouring and the use of a properly made 
babbitt metal, while other parts of the liner will not be in good con¬ 
tact with the shell, by reason of shrinkage and the formation of air 



bubbles and oxide gases, which latter are especially liable to be formed 
in babbitts containing copper. With the idea of filling up the hollow 
spaces betwe 3 n liner and shell, it is a quite general American practice, 
and an English one also, to peen or hammer the babbitt liner. “The 
advisability of this treatment is, however, very questionable. By the 
peening process the air will simply be driven from one point to an¬ 
other, and be forced into places where at first a good contact existed, 
thus destroying it. To secure a permanent and intimate contact be¬ 
tween liner and shell by peening is impossible, on account of the elas¬ 
ticity of the liner material. When the hammer strikes the metal, a 
contact may be formed, but as soon as the force of the blow is gone, 
the metal will spring away more or less by reason of its elasticity! 
Furthermore, the babbitt metal becomes more brittle by peening, and 
its strength diminished; this has been proved beyond doubt by a 
number of tests. Peening, unless performed with the utmost pre¬ 
caution, also produces minute cracks in the structure of the babbitt, 
which will constantly be enlarged by the regular working pressures! 
For these reasons, European continental practice has now practically 
abandoned the peening of babbitt metal liners. Summing up, in spite of 












19 


CAUSES OF HOT BEARINGS 

good pouring, or peening, or dove-tailed grooves and other similar 
anchoring devices, the liners are in a greater or less degree loose in 
the shells. 

The Lubricant Penetrating the Hollow Spaces 
When these loose bearings are in service, the hollow spaces between 
the liner and shell gradually become impregnated with an oil film 
from the lubricant employed, as shown in Fig. 12. Now, the coeffi¬ 
cient of heat-conductivity of oil is only about 1/200 of that of an 
ordinary babbitt metal, or of cast iron. Therefore, the heat created 
in the liner by the working friction will not be conducted away to the 
shell, and thence to the engine frame, as quickly as though an inti¬ 
mate contact existed between shell and liner. The result is that the 



Fig. 12. Penetration of Oil between Shell and Liner 

bearing readily becomes hot, because the babbitt metal liner retains, 
instead of throwing off, the heat. The regular working pressure also 
sets up a hydraulic pressure in the oil film, between the shell and 
the liner, which tends to produce breakages and cracks in the liner, as 
may sometimes be observed when removing bearings from gas en¬ 
gines, pumping engines and the like, subject to high pressures and 
shocks. A consequence of shocks is also that a liner which is some¬ 
what loose will become distorted and “work”; this “working” pro¬ 
duces additional friction and increased temperatures. All the facts 
mentioned above tend toward the one result, viz., the increasing of 
the temperature in the bearings, even to the extent of melting down 
the babbitt liner. 

From various tests which have been made, the results of one may be 





















20 


No. ii—BEARINGS 


given here. A bearing with a perfect contact between liner and shell 
was tested under a constant load of 400 pounds per square inch and a 
constant sliding speed of 480 feet per minute. The same bearing was 
again tested under the same conditions, but with the liner not in inti¬ 
mate contact with the shell. As the tests were necessarily made under 
a slightly varying atmospheric temperature, the difference between the 
actual hearing temperature and the room temperature was taken as 
the basis of each, and in the former case the result was 60 degrees 
F., while in the latter 85 degrees F. When such differences are ob¬ 
tained in a testing machine, under the best operating conditions, how 
much worse must be the influence of the slightest lack of contact 
under usual working conditions, such as we have them in steam en¬ 
gines, air compressors, pumps, gas engines, etc.! 

Summing up the foregoing we may say that in mo&t cases the direct 
causes of hot bearings are: A lack of contact between liner and shell, 
caused, first, by shrinkage and careless treatment of the babbitt, and 
second, by shrinkage strains produced by an unequal distribution of 
the liner masses over the shell; the formation of an isolating oil film, 
together with its consequences; cracks or breakages in the liner pro¬ 
duced as explained. The means of avoiding these troubles, and the 
principles of a good and safe bearing construction, must consequently 
be an absolutely intimate and homogeneous contact between liner and 
shell; an equal distribution of the liner over the shell; and a strength¬ 
ening of the liner against the shocks and working pressures. If 
these conditions are faithfully carried out, many troubles and much 
expense may be avoided. 


CHAPTER III 


THRUST BEARINGS 

Thrust bearings are, in general, of two kinds: step bearings and col¬ 
lar bearings. In the former the thrust is taken by the end of the sup¬ 
porting shaft, in the latter by projections or shoulders at some distance 
from the end of the shaft. The simplest kind of a thrust bearing is the 
pivot bearing, exemplified by the bearings for watch pinions and by a 
lathe center taking the end thrust of a cut on a piece held between the 
centers in a lathe. In general, however, the end thrust is taken by 

a b 



a large flat or nearly flat surface. When this is the case several con¬ 
siderations present themselves which must be given due attention by 
the machine designer. 

Assume that the flat end of a vertical cylindrical shaft carrying a 
weight or otherwise subjected to pressure is supported by a flat sur¬ 
face. Then, if the shaft rotates, the velocities of points on its end 
surface at different radial distances from its axis, will vary. The 
velocities of the points near the outside will be, in comparison, very 
high, while the velocity of a point near the center will be low. On 
account of this variation in velocity, the wear on the end surface of the 
shaft and the thrust surface of the bearing will be considerably un¬ 
even. If the parts are well fitted together when new, so that a uni- 






22 


No. ii—BEARINGS 


form pressure is produced all over the end of the shaft and bearing, 
then the outer parts of the bearing surfaces will wear away most 
rapidly. This again increases the pressure at the center, which some¬ 
times may become so intense as to exceed the ultimate crushing strength 
of the material. The unequal wear of the surfaces of thrust bearings 
is one of the most difficult problems meeting the designer of machinery 
of which such bearings form a part. 

Experiments carried out by Schiele show that the wear is theo¬ 
retically along a curve called the tractrix, the construction of which 
will immediately be referred to. If an end thrust bearing is made of 
a form corresponding to the Schiele curve, then the wear in the direc¬ 
tion of the axis of the thrust shaft will be uniform at all points; but 
while this curved form would be theoretically correct, it has been 
shown in practice that nothing is to be gained by the use of bearings 
having this complicated shape. 

The tractrix or Schiele curve may be constructed as follows. In 




Figs. 14 and 15. Simple Designs of Step Bearings 

Fig. 13 draw the lines A B and C D, the first representing the extreme 
diameter of the shaft or spindle and the latter its axis. Set off on A B 
a number of equal spaces, 1, 2, 3, 4, 5, etc., and on CD other equal 
spaces numbered to correspond and to suit the desired length of the 
bearing. Then join these spaces by the lines 1-1, 2-2, 3-3, etc. The in¬ 
tersections E will be in the path of the curve to be constructed. 

Simple Step Bearings for Light Duty 

For light duty simple step bearings of the types shown in Figs. 14 
to 17 answer the requirements well. The intense pressure at the 
center and the consequent unequal wear are partly avoided in the bear¬ 
ing in Fig. 14, by cutting away the metal at the center of the shaft, 
as shown, leaving an annular ring which takes the thrust. This pro¬ 
cedure is advisable in all step bearings. Another difficulty met with in 
bearings of this type is the question of lubrication. If the speed of 
the shaft is high, the centrifugal force tends to throw the oil out from 
the center. Special provisions must then be made for again returning 
the oil to the center, as otherwise the bearing would wear down rapidly, 







































THRUST BEARINGS 


23 


become heated, etc. In Fig. 14 a simple method is shown for auto¬ 
matically returning the oil to the bearing surfaces. An oil-passage is 
made from the chamber A, formed around the shaft, to the center of 
the shaft at the bottom. When the channel and chamber are once 
filled with oil, this oil will continue to circulate automatically; it 
will be drawn in at the bottom, be thrown outward by the centrifugal 
force, find its way into the chamber A, and finally, through the chan¬ 
nel, return to the center of the bearing. 

When a bearing for heavier duty is required, the design shown in 
Fig. 15 is quite commonly adopted. Here a number of disks or washers 
are placed between the end of the thrust shaft and the supporting 
bearing. The object of this is to introduce a number of wearing sur¬ 
faces, instead of having the end of the shaft and the box take all the 
wear. Due to the fact that the series of washers introduced permits 
of a lower speed between each pair of washers, the wear is quite ma¬ 
terially reduced. Should the pressure cause any two washers to heat 



Figs 10 and 17. Improved Designs of Step Bearings 

and bind, the frictional resistance between them ceases, as one washer 
is free to follow the motion of the other, and the oil will have an 
opportunity to get between the surfaces and cool them off. 

A hole may be, and generally is, drilled through the centers of the 
washers, as shown in Fig. 15, and the same method for continual 
lubrication, as shown in Fig. 14, may be used to advantage. Every 
alternate washer is commonly made of hardened tool steel or case- 
hardened machine steel, while the others are made of bronze. This com¬ 
bination provides for good wearing qualities. If the thrust shaft is 
made of soft machine steel, and the box of cast iron, the top washer 
is often secured to the shaft, and the bottom washer to the box, so that 
all the wear may be concentrated upon the washers, which can easily 
be replaced. 

In Fig. 16 is shown an improvement on the bearing in Fig. 15. This 
construction is recommended, in particular, in cases where the shaft 
and its bearing box cannot be properly aligned with one another. The 
washers have spherical faces, being alternately convex and concave. 
They are slightly smaller in diameter than the bearing box into which 







































24 


No. n—BEARINGS 


they are inserted, so that they may have an opportunity to adjust 
themselves to a perfect bearing on each other, and thereby make up 
for the differences in the alignment of the thrust shaft and bearing box. 

Another type of thrust bearing for loads which are not excessive is 
shown in Fig. 17. It is a well-established principle that it is better to 
take the thrust of a bearing as near the center of the shaft as the 
load to be carried will allow. The farther away from the center the 
support is, the greater is the motion and the greater is the retarding 
effect of the friction. The thin convex washers used are of tool steel, 
hardened, and although the bearing between them is very small, their 
strength and hardness is such that they are capable of standing a con¬ 
siderable pressure, though not as great a one, probably, as the other 
forms shown in Figs. 14, 15 and 16. In this bearing, also, there is no 



difficulty in keeping the surfaces well oiled, since all that is necessary 
is to keep the chamber well flooded with oil. 

Collar Thrust Bearings 

When a considerable thrust is to be taken care of, or when the 
thrust is taken on the shaft at a distance from its end, collar thrust 
bearings are used. They are usually of the form shown in Fig. 18. In 
a well made bearing each of the collar surfaces takes its proportionate 
part of the load, and it is thus possible, without using excessive diam¬ 
eters, to properly distribute a very great thrust on a number of collars 
formed solidly with the shaft by cutting a number of grooves in the 
latter. One advantage of the collar bearing is that the difference be¬ 
tween the outer and inner diameters of the bearing surface is not very 
great, and hence the velocities at the outer and inner edges do not 
vary appreciably; this, again, eliminates unequal wear on the thrust 
collar surfaces. 

































CHAPTER IV 


FRICTION AND LUBRICATION* 


Probably the most important and complete series of experiments on 
the friction of journals and pivot bearings yet undertaken was carried 
out by the late Mr. Beauchamp Tower for a Research Committee of 
the British Institution of Mechanical Engineers. In carrying out the 
experiments, as the result of an accidental discovery, an attempt was 
made to measure the pressure at different points of the bearing. A 
hole had been drilled through the cap and brass for an ordinary lubri¬ 
cator, when, on restarting the machine, oil was found to rise through 
the hole, flowing over the top of the cap. The hole was then stopped 
with a wooden plug, but this was gradually forced out on account of 
the great pressure to which the oil was subjected, and which on screw, 
ing a pressure gage into the hole was found to exceed 200 pounds 
per square inch, although the mean load on the journal was only 100 
pounds per square inch. Mr. Tower proved by this and subsequent 
experiments that the brass was actually floating on the film of oil 
existing between the shafting and the bearing. By drilling a number 
of small holes at different points in the brass, and connecting each 
one of them during the test to a pressure gage, Mr. Tower was able 
to obtain a diagram showing the distribution of pressure upon the 
bearing. It appears that the pressure is greatest a little to the off 
side and at the middle of the length of the bearing, gradually falling 
to zero at each edge. The total upward pressure was found to be prac¬ 
tically the same as the total load on the bearing, again showing that 
the whole of the weight was borne by the film of oil. Any arrangement 
w r hich would permit the film to escape was found to result in undue 
heating, and the bearing would finally seize at a very moderate load. 
The oil bath lubrication was found to be the most perfect system of 
lubrication possible. In the table below the results obtained by Mr. 
Tower are specified for three different methods of oiling. 


Actual Load in 
Pounds per 
Square Inch 

Oil bath. 263 

Syphon lubricator ... 252 

Pad under journal.... 272 


Coefficient 
of Friction 

0.00139 

0.00980 

0.00900 


Relative 

Friction 

1.00 

7.06 

6.48 


With the needle lubricator and a straight groove in the middle of 
the brass for distributing the oil, the bearing would not run cool when 
loaded with only 100 pounds per square inch, and no oil would pass 
down from the lubricator. The groove, in fact, w r as found to be a 
most effective method of collecting and removing the film of oil. In 
the next place, the arrangement of grooves usual in locomotive axle 
boxes was adopted, the oil being introduced through two holes, one 
near each end and each communicating with a curved groove. This 


* Machinery. March, 1907, and April, 1908. 




26 


No. ii—BEARINGS 


bearing refused to take the oil, and could not be made to run cool, 
and after several trials the best results which could be obtained led 
to the seizure of the brass under a load of only 200 pounds per square 
inch. These experiments proved clearly the futility of attempting to 
introduce the lubricant at that part of the bearing. A pad placed in 
a box full of oil was therefore fixed below the journal, so as to be 
always in contact with it when revolving. A pressure of 550 pounds 
per square inch could then be carried without seizing, or very nearly 
the same load as in the case of oil-bath lubrication. 


Results of Tower’s Experiments 

One important result was to show that friction is nearly constant 
under all loads within ordinary limits, and that it does not increase 
in direct proportion to the load according to the ordinary laws of 
friction. This is indicated by the result of the experiments recorded 
below. 

Variation of Friction with Pressure. —Journal, 4 inches diameter, 6 
inches long. Brass, 4 inches wide. Speed, 300 revolutions = 314 feet 
per minute. Temperature, 90 degrees F. 


BATH OP LARD OIL 

Pressure in pounds per square inch of bearing p 


W 

d X l 


Pressure per 

Coefficient of Friction 

Product 

sq. in. 

= A* 

P X 

520 

0.0013 

0.676 

415 

0.0016 

0.664 

310 

0.0022 

0.682 

205 

0.0031 

0.635 

153 

0.0041 

0.627 

100 

0.0067 

BATH OP OLIVE OIL 

0.670 

W 

Pressure in pounds per square inch of bearing p 

d X 

Pressure per 

Coefficient of Friction 

Product 

sq.in. 

= M 

PXp. 

520 

0.0013 

0.676 

468 

0.0015 

0.702 

415 

0.0017 

0.705 

363 

0.0019 

0.689 

310 

0.0021 

0.651 

258 

0.0025 

0.645 

205 

0.0030 

0.615 

153 

0.0044 

0.673 

100 

0.0069 

0.690 


The coefficient of friction with bath lubrication varies inversely as 
the pressure, or, in other words, the friction of the bearing is alto¬ 
gether independent of the pressure upon it; the first law of friction 
should therefore read: Temperature and velocity remaining constant 
the friction coefficient is proportional to the nominal pressure , and the 
work done against friction is independent of the load , provided this 
does not exceed from 400 pounds to 600 pounds per square inch. 
From this it follows that the work done in overcoming friction is 




FRICTION AND LUBRICATION 27 

independent of the load upon a machine, and that there is no appre¬ 
ciable increase in the loss due to friction from no load to full load. 
Under a load of 300 pounds per square inch and with a surface speed 
of 300 feet per minute, Mr. Tower found the coefficient of friction to 
be 0.0016 for oil-bath lubrication, and 0.0097 for a pad. 

In the next place it was found that the coefficient of friction is in¬ 
versely proportional to the temperature, other conditions remaining 
the same, as shown below. 

Variation of Friction with Temperature. —Journal, 4 inches diameter, 
6 inches long. Brass, 4 inches wide. Speed, 300 revolutions = 314 
feet per minute. Load, 100 pounds per square inch of nominal area. 


BATH OF LARD OIL 


Temperature 

Coefficient of 

Product 

Degs. F. 

(Degs. F. - 32) = t 

Friction = n 

tX M 

120 

88 

0.0044 

0.387 

110 

78 

0.0050 

0.390 

100 

68 

0.0058 

0.394 

90 

58 

0.0069 

0.400 

80 

48 

0.0083 

0.398 

70 

38 

0.0103 

0.391 

60 

28 

0.0130 

0.364 


The second law of friction should therefore be stated: Nominal 
pressure and velocity remaining constant, the coefficient, and therefore 
the work done against friction, is inversely proportional to the terns 
perature of the bearing. 

This has also been very neatly demonstrated by a recent experi¬ 
menter, Mr. Dettmar, whose machine is electrically driven, and there¬ 
fore the consumption of current could be very accurately measured 
during a five hours’ run at constant speed and voltage. As load and 
velocity remain constant throughout the test, a decrease in the loss 
due to friction could only occur with a diminution in the coefficient. 
The current fell off in the same ratio as the temperature increased. 

The results of Tower’s experiments seem to indicate that friction 
increases with the velocity, although not nearly in proportion to the 
square of the velocity as observed by Dettmar. As the result of the 
more exact determination possible with his machine, Dettmar found 
that friction increases very nearly as the 1.5 power of the velocity. 

The mean values of the coefficient of friction for different lubricants, 
and with different methods of lubrication as observed by Mr. Tower, 
are given in the following table: 

Variation of Friction with Different Lubricants. —Journal, 4 inches 
diameter, 6 inches long. Brass, 4 inches wide. Speed, 300 revolutions 
= 314 feet per minute. Temperature, 90 degrees F. 

Max. Safe Pressure in 
Coefficient of Pounds per«q. Inch 


Lubricant Friction of Nominal Area 

Olive oil. 0.00172 520 

Lard oil. 0.00172 570 

Sperm oil. 0.00208 570 

Mineral oil . 0.00176' 625 

Mineral grease . 0.00233 625 







28 


No. ii—BEARINGS 

Nicolson’s Experiments on Friction and Lubrication 

The remainder of this chapter consists of an abstract of a paper 
read by Dr. J. T. Nicolson before the Manchester (England) Associa¬ 
tion of Engineers. The chief aim of the Nicolson experiments was to 
give some definite ideas about the resistance offered to the relative 
motion of lubricated surfaces, and they, in particular, related to jour¬ 
nals and bearings as used in engineering practice. Experimental 
results obtained by Stribeck, Dettmar, Heimann, Lasche, and others 
have been utilized for framing rules which indicate that some views 
commonly held in regard to bearings are not correct. In particular, 
the idea that the length of the bearing should increase in proportion 
to the speed is shown to be erroneous. 

Dry Friction 

When one solid rubs upon another without any lubricant, the resis¬ 
tance offered to relative motion is due either to actual abrasion or to 
molecular interference between the two surfaces. Even though a 
metallic surface may appear to be perfectly smooth to the eye, its real 
condition, if viewed with a powerful microscope, resembles that of a 
rugged mountain system. When one surface is slid upon another, 
these surfaces exercise a resisting force. The following laws may be 
considered as generally covering the question of dry friction: 

1. Within certain limits, the frictional resistance may be said to 
be proportional to the load, and to be independent of the extent of the 
surface over which the load is distributed; but when the pressure or 
load per unit area is large, the friction inceases at a greater rate 
than the load, or, in other words, the coefficient of friction increases 
with the pressure. 

2. The coefficient of friction varies with the speed of motion. It is 
greatest when the motion is slowest, and when one body is just com¬ 
mencing to move relative to another, we have what is called friction 
of repose. This friction has been found by experiments to be from 
0.3 to 0.4 for iron upon iron; for moderate speeds the friction varies 
from 0.15 to 0.25 for the same material; and for speeds from 10 to 90 
feet per second, coefficients of from 0.10 to 0.20 have been found 
by experiments. 

3. The friction of solids with no lubricant interposed has been 
found to diminish as the temperature increases. This is due to the 
fact that abrasion is easier at high temperatures. 

Friction and Lubrication 

When some lubricant is placed between moving bodies, the valleys 
or the uneven surfaces are leveled up, and the intensity of the molecular 
action is diminished. For the frictional work when a shaft rotates 
in a well lubricated bearing, we may state the following formula, ex¬ 
pressing the frictional work done per revolution: 

ir d tlW 


Frictional work per revolution = 


12 


foot-pounds. 



FRICTION AND LUBRICATION 


29 


In this formula, 

d = diameter of shaft in inches, 
m = coefficient of friction (0.15 on an average), 

W = load on the bearing in pounds. 

This formula holds true when there is plenty of oil, so long as the 
speed is small. If we take as an example the case of the spindle for 
a 10-inch lathe, running slowly, with a weight of 3000 pounds carried 
by the front bearing, which is 3 y 2 inches in diameter, then the friction 
work per revolution is 

7T X 3.5 X 0.15 X 3000 

----—= 412 foot-pounds per revolution. 

12 

If a cut were 1/4 inch X 1/16 inch on soft steel, the cutting force 
would he, say, 3500 pounds, and on a 20-inch face-plate diameter the 
work spent in cutting per revolution would be 

20 7T 

3500 X-= 18,300 foot-pounds. 

12 

The work lost in friction by the journal is therefore 2.26 per cent 
of the useful work. A similar calculation for a 48-inch lathe would 
show a loss of about 10 per cent. These great frictional losses con¬ 
stantly occur with lathe spindles or other rotating shafts, revolving 
slowly, even when abundantly fed with oil, and indicate the necessity 
for using measures to preserve a separating film of oil between the 
shaft and bearing, and not to allow them to run in metallic contact. 
This is more difficult to accomplish at slow than at high speeds. 

Automatic Lubrication 

The following rules for supplying bearings with oil will give the 
best results in practice: If the oil is fed in by the ordinary cup and 
syphon, or by a ring or centrifugal method of supply, it should be 
made to flow onto the journal at the place where the pressure is least. 
The oil should therefore be fed from a point situated in the top rear 
quadrant of the bearing when the journal is loaded by gravity only, 
and the point should be further back the slower the speed. This 
applies, then, especially to the large lathes. If the loading of the 
journal is principally due to cutting force acting upward upon it, the 
feed should be placed in the bottom front quadrant, and nearer the 
front, the slower the speed of rotation. This meets the case of the 
smaller sized lathes. 

The compromise ordinarily effected to enable the lubricant to enter, 
whatever may be the direction of the loading, is the simple one of 
fitting the oil cup on the top of the bearing. This seems almost the 
only thing to do in the case of automatic lubrication, but it is the 
correct position only when the resultant force upon the journal, due 
to gravity and cutting force, etc., acts nearly horizontally and from 
front to rear. 

Forced Lubrication 

When the lubricant is supplied by mechanical means at a fixed rate 
and at any required pressure, it must be fed in at the points of greatest 




30 


No. ii—BEARINGS 


oil pressure in the bearing. For large lathes, where gravity is more 
important, the region of greatest pressure lies in the rear bottom 
quadrant. For small lathes, on the other hand, in which the force on 
the spindle acts upward, owing to the cutting force being relatively 
greater, the maximum oil pressures occur in the front top quadrant. 
To meet all contingencies, it would appear on the whole best, in the 
case of forced lubrication, either to force the oil in at the back of the 
bearing, well below the center, or preferably to fit three alternative 
branches from the oil pressure supply pipe to the back, top, and front, 
any one of which may be turned on at will to suit the conditions of 
working. 

Frictional Resistance Due to Viscosity 

In describing the phenomena occurring when a journal rotates in 
a bearing, we have, so far, not alluded to the nature or magnitude of 
the' frictional resistance experienced when there is an abundant supply 
of lubricant completely separating the former from the latter, and 
preventing any metal-to-metal contact. It is frequently stated that 
“there is no friction without abrasion,” or, in other words, that unless 
two metals rub against each other there can be no resistance due to 
relative motion. This, however, is not the case. When a film of lubri-. 
cant is interposed between two metallic surfaces there is a resistance 
to relative motion of these surfaces due to the shearing or transverse 
distortion of the oil film. 

This resistance does not depend on the load. It is governed only 
by the area of viscous fluid to be sheared and the viscosity of the oil, 
i. e., the kind of oil and its temperature (with which the viscosity 
greatly alters), and it also gets greater the smaller the thickness of 
the film, so that if the shaft is a close fit within its bearing the resist¬ 
ance to motion will be greater than if the fit is an easy one. 

There are very few cases in engineering practice where a journal 
rotates with a uniform thickness of oil around it, and it is only at 
very high speeds that this takes place. At moderate and low speeds 
the shaft moves to one side an amount depending on the speed of the 
load, the eccentricity for any given load becoming less the greater the 
speed. We have already said that the frictional resistance depends 
on the thickness of the oil film. Experiments have shown, however, 
that the thickening of the film on one side of the shaft is more than 
counteracted by the thinning of the film on the other, so that, in gen¬ 
eral, the friction gets greater when the journal becomes more eccentric. 

Considering, therefore, the bearing running slowly, in which a lubri¬ 
cant has just formed a complete film all around the shaft, it will have 
its maximum amount of eccentricity, and the frictional resistance 
will, on this account, be large. As the speed increases, the eccentricity 
diminishes. The friction increases with the speed, but it diminishes, 
on the other hand, with the eccentricity. Experiments show that at 
first there is a decrease and then an increase, so that the coefficient 
of friction attains a minimum value which depends on the circum¬ 
stances in each case. With further increase in speed, the diminishing 


FRICTION AND LUBRICATION 31 

of friction, due to the lessening eccentricity, becomes insignificant, and 
after a certain interval the simple law of friction is followed, whereby 
friction increases in proportion to the velocity of rubbing. 

For speeds greater than at from 20 to 80 feet per minute, the tem¬ 
perature of the oil film also exerts its influence. This temperature 
rises above that of the bearing, and its viscosity becomes reduced. 
The frictional resistance then increases less rapidly than in exact pro¬ 
portion to the speed. The faster the journal runs, the more the tem¬ 
perature of the oil film rises above that of the bearing, and the thinner 
or less viscous becomes the oil. Thus, for speeds from 50 to 90 up to 
about 450 feet per minute, the coefficient of friction is proportional to 
the square root of the speed of rubbing. For speeds between 450 feet 
and 800 feet per minute the friction increases more slowly, and varies 
as the fifth root of the velocity. For speeds as high as 3,600 feet per 
minute and upward, the influence of the speed disappears altogether, 
and the conclusion is arrived at that for bearings of high-speed gen¬ 
erators, for instance, driven by steam turbines, whose rubbing speeds 
are nearly a mile a minute, the coefficient of friction is the same, what¬ 
ever be the speed. 

Application of Results of Experiments to the 
Design of Bearings 

In endeavoring to apply the theoretical explanations and the experi¬ 
mentally found formulas, the question arises: What is the proper 
proportion of length to diameter, under any given condition, as to load, 
speed and kind of lubrication? According to hitherto accepted rules, 
the length of the bearing should increase with the load and with the 
number of revolutions. The experiments and formulas arrived at by 
the author indicate, however, that the heat developed in the bearing 
depends only upon the rubbing velocity, and is quite independent of 
the length of the journal. We cannot, therefore, hope to lower the 
temperature by lengthening the bearing. The heat generated increases 
as fast as the area for dissipating it increases, and, although by length¬ 
ening the journal the bearing pressure is diminished, the frictional 
resistance and the heat generated are increased. On the other hand, 
we know from experience that journals must be made long for 
high speeds, and the above calculations seem, at first sight, to be in 
conflict with accepted practice. The explanation of this is as follows: 
While it is true that the final temperature to which the bearing will 
rise after a long run, under a given load, and with a given lubricant, 
depends only on the diameter of the spindle and the speed of revolu¬ 
tion, that is, only upon the rubbing velocity, and not at all upon the 
length of the journal, we have to remember that if the finally attained 
temperature be too high, the lubricant will be squeezed out unless the 
bearing pressure is low. 

Another conclusion arrived at by these experiments, contrary to the 
view usually accepted, is that the length of the bearing must be 
greater, the slower the speed. This, however, is clearly correct, for 
the slower the speed, the greater difficulty has the shaft in dragging 


32 


No. ii—BEARINGS 


in its supply of oil to meet the required demand, in opposition to the 
bearing pressure which is squeezing it out, and consequently the unit 
bearing pressure should accordingly be lower in order to enable the 
journal to maintain its oil film unbroken. 

Journals for Heavy Loads at Slow Speed 

One kind of bearing which presents special conditions, and which 
is frequently met with and has to be dealt with in practice, is that in 
which a journal has to run under a heavy load at a very slow speed. 
What we have here to guard against is the entire collapse or tearing 
asunder of the film of lubricant, owing to the slow speed at which the 
bearing is being worked; and when once the tearing of the oil film 
begins, the journal is unable to bring up a fresh supply, owing to its 
small surface speed. 

Calculations and experiments show that it is impossible to give the 
large dimensions to the front bearing of a heavy lathe that would be 
necessary to prevent the oil film from being broken at such slow speeds; 
and, as a matter of fact, lathe spindles turning at the slow speeds 
used for heavy cuts inevitably run metal-to-metal with their bearings, 
giving rise to the high frictional resistance corresponding to the co¬ 
efficient of friction of 0.15 for greasy metals. The work thus spent 
and wasted on friction and wear may amount to from 2 per cent to 10 
per cent of the total useful work expended on cutting. From % to 9 
(according to size) horsepower is, therefore, wasted on the friction of 
the front journal alone when the lathe is running at these slow rates 
with a heavy job between centers. Even if the working pressure is 
light, and the thrust on the front journal is due to the standard cut 
only, it can be shown that 2% per cent of the useful work is spent on 
friction on any size of lathe when the speeds are so low as to squeeze 
out the oil film. 

We are here face to face with a very serious loss of power, and a 
correspondingly large amount of wear of the spindle and in the front 
bearing, not at all due to high speeds of rotation of the spindle; and it 
is owing to this that the elaborate arrangements for adjustment of 
the spindle in a lathe head-stock have to be provided. 

It is impossible to give enough area in the front bearing of a lathe 
head-stock to prevent metallic contact of journal and brass at the 
slower speeds, if dependence is placed upon the lubricant being carried 
in by the ordinary action of the shaft’s rotation, the supply being auto¬ 
matic. By using a force pump, however, and injecting a stream of 
moderately heavy oil into the bearing at the place where the pressure 
is greatest, it is possible to raise the journal off the brass even when 
at rest, and to keep it floating with a film of oil interposed between 
itself and the bearing when in motion, be that motion as slow and the 
load as high as it may. If metal-to-metal contact can in this way be 
prevented at slow, and by the ordinary methods at high speeds, there 
seems to be a possibility that wear may be entirely eliminated. If 
this be so, it follows that adjustments for wear are unnecessary, and 
instead of the elaborate and expensive designs of front and back bear- 


FRICTION AND LUBRICATION 33 

ing which are now used, we may expect that a simple solid bush of 
ample thickness will meet every requirement. Such a solid bush, of 
hard bronze round the steel spindle, has a great deal to recommend it 
from the point of view of accuracy of fit, solidity, and stiffness, as 
compared with the intricate methods of adjustments now common. 

Modern Practice for Lubricating- Bearings 

The chief distinction between the modern and the older methods of 
lubricating bearings lies in that the oil is no longer supplied drop by 
drop, as formerly, but in an abundant stream, the oil serving the pur¬ 
pose not only of lubrication, but of carrying away the heat. 

For high speed bearings, the principle most often adopted is that 
of the “closed circuit”; that is, the oil is used over and over again; 
after dropping off the journal into a collecting reservoir it is filtered 
and used anew, being automatically supplied to the journal at any 
suitable point, A cooling arrangement is sometimes fitted in the reser¬ 
voir, so as to remove the heat from the oil, and consequently also 
from the bearings. The system of forced lubrication is also adopted 
to a great extent. The oil is then, by means of a pump or other suitable 
device, pressed in between the rubbing surfaces so that the journal 
floats on the heavy film of lubricant. 

Lubricating Horizontal Bearings 

The most common method of lubrication for horizontal journals run¬ 
ning at high speed is the ring-oiled bearing, in which a loose ring, 
resting on the shaft, turns with it, dipping into the oil reservoir at 
the lower side, and bringing up the oil to the top surfaces of the jour¬ 
nal, from where it flows over into the oil grooves. No ribs or other 
projections should be fitted on the rings, as such arrangements produce 
a resistance to their passage through the oil bath, and bring them to a 
standstill. At high speeds, the centrifugal force renders the flow of 
oil from the ring to the journal difficult, and scrapers are used for 
diverting the oil into the oil channels. These, however, should never 
touch the ring, as they will then stop its motion. 

Self-oiling bearings having rings fast on the shaft are not much used. 
The fast ring cannot stick, but it requires a longer design of bearing. 
The ring may act as a collar where endwise motion is to be prevented; 
but as such motion is usually an advantage, the ring should ordinarily 
be attached to the shaft so that it can slide on its key. For high 
speeds the scraper may be used with fast rings, to overcome the cen¬ 
trifugal force. 

Forced Lubrication 

By the use of a pump to force the oil drawn from the reservoir into 
the bearing to the point of maximum pressure, the length of the bear¬ 
ing can be very much diminished even for the slowest speeds, espe¬ 
cially for journals whose load and rotation direction do not change. 
For such bearings the length need, in all probability, not be more than 
equal to the diameter of the shaft. With such bearings there ought 


34 


No. n—BEARINGS 


hardly to be any wear at all. The system is extensively used in high 
speed steam engines and gas engines. 

Grease as a Lubricant 

Grease has certain advantages as a lubricant which make its use 
advisable in many places, but it should not he expected that its lubri¬ 
cating value is ever as good as that of the best oil, although it may 
give better results in some places. For example, grease is particularly 
valuable for bearings exposed to dust, for when it is forced into the 
bearings with compression grease cups, the grease flows outward 
around the journals, forming a perfect dust protector, both because it 
seals the bearing and because the outward flow of the grease repels 
the intrusion of dust and abrasive particles. In such places the best 
oil would not give nearly as good results as grease, although its 
lubricating quality is generally considerably greater. On the other hand, 
the use of grease for lubricating machinery of a mill would not be 
advisable where the power factor is important in the cost of produc¬ 
tion. For example, some tests w*ere made several years ago in the 
lubrication of the machinery of a flour mill that was run by two water 
wheels of the same size, as stated by Mr. W. F. Parish, Jr., in a paper 
read before the North Eastern Coast Institute of Engineers and Ship 
Builders. In making the trial of grease the section driven by No. 1 
w^ater wheel w r as fitted up first. As the grease displaced the oil it was 
noticed that the speed of the mill decreased with a consequent de¬ 
crease of production. At first no one thought that the grease was 
responsible for the slowing down, but as the second part of the mill 
slowly decreased in speed as the use of the grease was extended, a con¬ 
sulting engineer was called in, who suggested that, in view of the fact 
that speed had decreased with the introduction of grease, it was respon¬ 
sible for the loss of production. Upon the resumption of the use of 
oil the speed of the machinery again rose to its original figure, proving 
conclusively that the lubricating value of the grease was inferior to 
that of oil anq that the difference wa,s an important factor in the 
mill’s production. The relative value of different oils in the lubrication 
of textile mills has long been known to be important in influencing the 
cost of production. 


CHAPTER V 


BEARING METALS* 

By conservative estimate the value of the hearing metal in actual 
use in the United States exceeds $50,000,000, of which fully one-half 
is used on the locomotives and rolling stock of the railroads of this 
country. In view of the increase in the amount of machinery and 
rolling stock steadily going on, and the constant wearing out and 
replacement of bearings, the value and importance of this product 
cannot be overestimated. The life of a machine is largely dependent 
upon its bearings, and in view of this the fact that knowledge in regard 
to bearing metals and alloys is not more general, is remarkable. Again, 
the nature of the production of these alloys is such that, while in 
some cases they have been patented and are manufactured under trade 
names, in many others they are made up of scrap, with widely varying 
proportions of the different metals incorporated in their structure; on 
this account, probably no phase of engineering progress in machinery 
construction and operation is the subject of more difficulty and dis¬ 
satisfaction. 

The fact that bearing metals have to be taken largely on faith or 
else tested by more or. less complicated processes for their chemical 
constituents, and the further fact that trade conditions in this field 
are such that the properties of metals are apt to vary greatly in dif¬ 
ferent shipments, is a matter of grave import to the average machinery 
manufacturer and operator. Only the largest consumers can afford to 
make the necessary tests and investigations of a given consignment 
in order to test its quality, and, in addition, a definite amount of spe¬ 
cial knowledge is requisite for this purpose, in view of the often 
wide variations in properties of the alloy, with a comparatively smaU 
variation in the proportion of its constituents. Under these circum¬ 
stances the average small machine shop and consumer in this field 
accepts bearings on faith alone and is dependent largely upon the 
commercial reputation of the firm furnishing the material. That this 
should not be so is a foregone conclusion, but in view of this condition 
of affairs the rapid progress of the firm whose standing can be relied 
upon in this field is readily explained,, 

Bearings are usually composed of alloys of copper, lead, tin, antimony 
and zinc, and are known as babbitt metal (after the name of the dis¬ 
coverer of this material), white metal, brass, phosphorous bronze, and 
various other trade names. Quite a number of these are patented, 
such as “plastic bronze,” etc., but many are sold merely under trade 
name, and in some instances are of uncertain composition. 

The principal qualities which a good bearing metal should have are 
good anti-frictional properties, so as to withstand heavy loads at high 
speed, without heating, and, second, sufficient compressive strength so 

♦Machinery, August, 1909. 




36 


No. ii—BEARINGS 


as to neither be squeezed out of place under high pressure, nor crack 
or break when subjected to sudden shocks. In addition to these, many 
other properties must be considered in a choice of bearing metals 
depending upon the special purpose for which the material is to be 
utilized. Temperature variation is often an important factor, espe¬ 
cially in refrigerating plants, and the coefficient of expansion should 
be considered to prevent undue binding, with consequent destruction of 
the bearing and the possible variation in other properties, such as 
brittleness, ductility, etc., under various temperature conditions. In 
addition, many bearings must operate under conditions where they are 
subject to chemical action, whether that of brine or ammonia in refrig¬ 
erating plants, or acids, alkalies, etc., in chemical establishments, and 
in dynamo and motor construction and operation, the electrical con¬ 
ductivity must be considered as well. This statement applies equally 
to all bearings incorporated in electrical machinery, where these must 
serve as electrical conductors such as the bearings for the wheels in 
trolley cars, etc. 

The chief properties to date which have been developed to a greater 
extent than others in machine design are those of friction elimination 
and resistance to compressive loads. Theoretically, all metals have 
the same friction, according to Thurston, and the value of the soft 
white alloys for bearings lies chiefly in their ready reduction to a 
smooth surface after any local impairment of the surface, such as 
would result from the introduction of foreign metal between the moving 
surface and the bearing. Under these circumstances the soft alloys 
flow or squeeze from the pressure into the irregularity, forming a larger 
area for the distribution of the pressure, thus diminishing its amount 
per unit of area. Further, the larger area over which the pressure is 
extended the less becomes the liability to overheating and consequent 
binding. Under these circumstances the frictional properties of a bear¬ 
ing are in inverse ratio to their compressive resistance, and invariably 
the best bearing alloys, from a high speed standpoint, are unsatisfac¬ 
tory for utilization in heavy machinery. The recent introduction of 
an iron or steel grid to form the base of the main bearing, and to be 
filled with much softer bearing metals than could ordinarily be in¬ 
stalled, or in some cases even graphite, is a step in the right direction 
and presents possibilities of great importance in this field of machine 
development. 

Lead flows more easily under pressure than any of the common met¬ 
als, and hence it has the greatest anti-frictional properties. Of course, 
a number of metals exceed lead in this property, but their cost or some 
other factor render them unavailable. Lead is the cheapest of the 
metals, except iron, and in comparison to the other metals used in the 
formation of bearing alloys their relative prices are somewhat in the 
following order per one hundred pounds: Lead, $4; zinc, $5; antimony, 
$9; copper, $13; and tin, $30 or more. It can thus be seen that the 
more lead that is used in a given bearing, the softer it is, the less fric¬ 
tion it possesses, and the cheaper it can be furnished. It is, however, 
too soft to be used alone, as it cannot be retained in the recesses of the 


BEARING METALS 37 

bearing even when used simply as a liner and run into a shell of brass, 
bronze or gun-metal or some other alloy. Various other metals have 
been alloyed with it, such as tin, antimony, copper, zinc, iron and a 
number of non-metallic compounds, such as sodium, phosphorus, car¬ 
bon, etc., and the effect of the different ingredients is to-day fairly 
well understood. 

If antimony is added to the lead it increases its hardness and brittle¬ 
ness, and if tin is added as well it makes a tougher alloy than lead or 
antimony alone. Nearly all of the various babbitt metals on the mar¬ 
ket are alloys of lead, tin and antimony in various proportions, with 
or without other ingredients added. In such babbitts the wear in¬ 
creases with the antimony as a general thing, and the price with the 


COMPOSITION OF BFARING METALS. 


Alloys 

Lead. 

Tin. 

Anti- 

mony. 

Cop¬ 

per. 

Zinc. 

Other Con¬ 
stituents. 

Babbitt 1.. 

80,00 

72.0 

70,0 

80.5 

0.5 

20.0 

21.0 

10.0 

11.5 

68.0 

20.0 

86.0 





Babbitt 2. 

7.0 

20.0 

7.5 




Babbitt 3. .. 




Babbitt 4... 

0.5 

1.0 


■ 

Babbitt 5. 

81.5 

80.0 


Babbitt 6. 



Babbitt 7... 


10.0 

12.0 

'ik'.o 

7.5 

4.0 

6.0 

2.00 

trace 

0.5 

65.0 

77.0 

77.0 

79.7 

92.4 

70,2 

0.11 


White metal. 

82,0 

80.00 

80.5 
30.0 

11.5 
15.0 

9 5 
6.1 
14.8 

98.5 



White Brass.. 

Magnolia metal.... 
Car brass lining 
Ajax plastic bronze. 

Ajax metal. 

P. R. R. car brass, B. 
S bearing metal , .. 

Delta metal- .. 

Oamelia metal. 

Tempered lead. 

64 0 
4.75 
11.5 
5.0 
11.5 
8.0 
10.0 
2.4 
4.3 
0.08 

34.0 

10.2 

Bi = 0.25 

”p = olio 

Fe = 0.1 
Fe- 0.5 
Na = 1.30 


JBi =; bismuth; P = phosphorus; Fe = iron; Na = sodium. 


tin. The higher antimony babbitts are used in heavy machinery, as 
they are harder, while those low in antimony are used in high speed 
machinery. The steady increase in speed at which various operating 
units are maintained is responsible for a wide deficiency in this field 
in the duty performed by the bearing metal. The chief difficulty to¬ 
day in the operation of the modern turbine is undoubtedly the main¬ 
tenance of satisfactory bearing surfaces. Soft babbitts have never 
sufficient strength to sustain the weight and shock of heavy machinery 
bearings and can only be used as liners. The tendency to increase in 
speed as well as weight or size of machinery is limited to-day simply 
by the satisfactory operation of the bearing metal itself. 

Undoubtedly, in investigations in this field, sufficient attention has 
not been paid to the effect of temperature on the bearing properties 
of the alloys used for these bearings. More rigid investigation in this 
field and limitations in regard to the temperatures permissible, with 


















































38 


No. ii—BEARINGS 


means for maintaining these within fairly close limits, will undoubtedly 
result in a great increase in the possibility of improvements in speed 
and weight of various types of machinery. More or less extensive 
experiments along these lines are being conducted in regard to the 
bearings used in turbine construction, since the speed here has ren¬ 
dered the problem an acute one and is necessary for efficient operation 
of the turbine itself. 

The accompanying table will doubtless prove interesting as showing 
the various constituents of the more or less common bearing metals 
now on the market. The original babbitt metals were very expensive 
materials, on account of the proportions of the more expensive metals 
found in them, and have been much modified in actual practice. A 
wide deviation in the composition of babbitt is readily shown in the 
first part of the table. The first babbitt is a fairly good alloy for high 
speed machinery, but is not very hard. Its melting point is about 500 
degrees F.; in fact, the properties of all allows or bearing metals can 
be very widely deduced from their melting point. The second babbitt 
is somewhat harder and melts at a higher point. Both of these are 
used largely for lining purposes. The fourth babbitt is used very 
widely for heavy machinery. All of the babbitts mentioned have been 
fairly successful. 

Babbitt 6 has good wearing properties, but cannot be used for high 
speeds. Most of the other metals included in the table where copper 
is not used in excess can be regarded as in the same class as babbitts. 
The “white” class has a fairly good electrical conductivity, much 
greater than that of ordinary babbitt, and is used in the bearings of 
generators, motors, electric cars, etc. A rather interesting thing about 
the alloys containing sodium is based upon the fact that sodium by 
oxidation produces a material which will saponify with the oil used 
in the bearing and produce soap, thus assisting lubrication. The ex¬ 
tent and amount of such action is scarcely as yet understood, and 
practically no experiments have been made with this investigation in 
view. Possibilities along this line, however, are great, not only for 
this particular alloy, but for many others not as yet considered. 

The other alloys included in the table consist to a very great extent 
of copper, tin, and lead, and usually have a thin liner of lead or some 
soft babbitt, and hence wear much better than an entire bearing of 
the soft babbitt. The tendency to wear decreases with increase of lead 
and increase of tin. Increase of lead, of course, affects the frictional 
quantities of the alloy and hence its heating properties. A certain 
amount of other metal, however, is necessary to keep the lead from 
separating from the copper. A study of the table itself, with a knowl¬ 
edge of the various properties of the metals themselves, will show 
conclusively the bearing properties of the different alloys. Pure copper 
is so tenacious that it is practically impossible to work it with any 
tools whatever without preliminary treatment, and this same property 
extends into and influences its bearing properties. 

The structure and treatment has more to do with the production of 
suitable bearing alloy than is generally considered. The tensile 


BEARING METALS 


39 


strength of solder and, in fact, all alloys, decreases very greatly with 
the pressure or tension at the time of solidification, and in general the 
cooling process, and the influence on tempering, affect the structure 
and consequently compressional resistance to a much greater extent 
than is generally considered. The same properties which influence 
the hardening and tempering of steel by heat, extend to a greater or 
less degree to all metals and are much more pronounced in alloys 
than in the simple elements. 

Sufficient has been said to show the importance of the bearing metals 
in machine design to-day, and to give a brief outline of the situation 
in regard to the character and type of the metals available, with a 
few of the properties of the same. The possible combinations of alloys 
for this purpose are very great. Comparatively little progress has been 
made along investigations covering all possible alloys of different ma¬ 
terials in different proportions. The recent introduction and placing 


COMPOSITION OP BRONZES 

White Metal: Parts. 

Tin . 7.6 

Copper . 2.3 

Zinc . 83.3 

Antimony . 3.8 

Lead . 3.0 

Hard Bronze for Piston Rings: 

Tin .. 22.0 

Copper . 78.0 

Bearings—Wearing Surfaces, etc.: 

Copper . 6 

Tin . 1 

Zinc . % 

Naval Brass: 

Copper . 62.0 

Tin . 1.0 

Zinc . 37.0 

Brazing Metal: 

Copper . 85.0 

Zinc . 15.0 

Anti-friction Metal: 

Copper—(best refined) . 3.7 

Banca tin . 88.8 

Regulus of antimony. 7.5 

Well fluxed with borax and rosin in mixing. 

Bearing Metal—(Pennsylvania Railroad): 

Copper . 77-0 

Tin . 8.0 

Lead . 15-0 


on the market of a large number of metals, such as calcium, etc., very 
common in nature, and ultimately bound to be furnished at a very low 
rate, and many of them possessing very suitable properties for bear¬ 
ing alloys, is undoubtedly bound to influence the situation; and various 
engineering devices, such as the steel grid, recently developed, will 
undoubtedly receive attention in the immediate future with consequent 
increase in efficiency in this field. The development is but at its incep¬ 
tion along this line, and standardization of the alloys at hand should 























40 


No. n—BEARINGS 


be at once insisted upon and maintained by the various machine 
manufacturers. This latter is the chief difficulty to-day in commercial 
development. The scientific end will largely take care of itself. The 
effect of different metals upon alloys by their presence in various pro¬ 
portions can be foretold to-day largely from theoretical considerations; 
but that the commercial situation to-day, however, is unsatisfactory, is 
a foregone conclusion. 

The table on the preceding page, giving the composition of bronzes 
used by the U. S. Navy Department, was contributed to Machinery’s 
Data Sheets by Mr. F. W. Armes, and is reproduced from Data Sheet 
No. 31, April, 1904. 


CHAPTER VI 

ALLOYS FOR BEARINGS* 

In an important article, in the Journal of the Franklin Institute for 
July, 1903, Mr. G. H. Clamer discussed the advantages and disadvan¬ 
tages of various compositions and alloys for bearings, and especially 
alloys for railway journal brasses. He also quoted the results of many 
tests on various compositions made on an Olsen testing machine de¬ 
signed by Prof. Carpenter of Cornell University. The present chapter 
is devoted to an abstract of Mr. Clamer’s article, and contains all the 
most important features of his discussion on a subject on which not 
so much is generally known as would be desirable. 

Upon close examination we find that there are but few metals avail¬ 
able for bearings. As mentioned in the previous chapter, they are 
copper, tin, lead, zinc and antimony. While other metals may be intro¬ 
duced in greater or less proportions, the five mentioned must constitute 
the basis for the so-called anti-friction alloys. The combinations of 
these metals now used may be grouped under the two heads of white 
metal and bronze. Bronze is the term which was originally applied 
to alloys of copper and tin as distinguished from alloys of copper and 
zinc; but gradually the term “bronze” has become applied to nearly 
all copper alloys containing not only tin, but lead, zinc, etc., and no 
sharp lines of demarcation exist between the two. 

Principal Requirements of Bearing Metals 

White metals are made up of various combinations of lead, anti¬ 
mony, tin, copper and zinc, and may contain as few as two elements, 
or all five. Bronzes are made up of combinations of copper, tin, lead 
and zinc, all of them containing copper and one or more of the other 
elements. The essential characteristics to be considered in any alloy 
for bearings are composition, structure, friction, temperature of run- 


* Machinery, October, 1903 



41 


ALLOYS FOR BEARINGS 

ning, wear on bearing, wear on journal, compressive strength, and cost. 

It is utterly impossible to have one alloy reach the pinnacle of per¬ 
fection in all of these requirements, and so it is important to study 
the possible compositions and determine for what purpose each is 
adapted. It has been shown that a bearing should be made up of at 
least two structural elements, one hard constituent to support the load, 
and one soft constituent to act as a plastic support for the harder 
grains. Generally speaking, the harder the surfaces in contact, the 
lower the coefficient of friction and the higher the pressure under 
which “gripment” takes place. It would seem for this reason that 
the harder the alloy the better; and it was with this idea in mind that 
the alloys of copper and tin were so extensively used in the early days 
of railroading. A hard, unyielding alloy for successful operation must, 
however, be in perfect adjustment, a state of affairs unattainable in 
the operation of rolling stock. For this reason the lead-lined bearing 
was introduced and the practice of lining bearings has now become 
almost universal in this country. 

General Comparison between Hard and Soft 
Alloys for Bearing's 

While the harder the metals in contact the less the friction, there 
will also be the greater liability of heating, because of the lack of plas¬ 
ticity, or ability to mold itself to conform to the shape of the journal. 
A hard, unyielding metal will cause the concentration of the load upon 
a few high spots, and so cause an abnormal pressure per square inch 
on such areas, and produce rapid abrasion and heating. 

The bronzes will, generally speaking, operate with less heat than 
softer compositions, while the softer metals will wear longer than the 
harder metals. In the matter of wear of journal, however, the soft 
metals are more destructive. Particles of grit and steel seem to be¬ 
come imbedded in the softer metal, causing it to act upon the harder 
metal of the journal like a lap. High-priced compositions are being 
used that have but little resistance to wear compared with cheaper 
compositions, and low-priced alloys are in service that are not cheap 
at any price. It is generally conceded that soft metal bearings cause a 
marked decrease in the life of the journal, and yet they have many 
marked advantages, as we shall presently see. 

Alloys Containing Antimony 

1. Lead and Antimony: These metals will alloy in any proportion. 
With increase in antimony the alloy becomes harder and more brittle. 
It has been determined that when it is made of 13 parts antimony 
and 87 parts lead, the composition will be of homogeneous structure. 
If there is a greater proportion of antimony, free crystals of antimony 
appear, imbedded in the composition; and if less than 13 per cent, 
there appear to be grains of the mixture itself imbedded in the lead 
as the body substance. 

According to one writer, an anti-frictional alloy should consist of 
hard grains, to carry the load, which are imbedded in a matrix of 
plastic material, to enable it to mold itself to the journal without undue 


42 


No. ii—BEARINGS 


heating. Such a condition would be met in a lead and antimony alloy 
having above 13 per cent antimony, but it is not advisable to use in 
any case more than 25 per cent antimony, as the composition would be 
too brittle. The same writer claims that alloys having from 15 to 25 
per cent antimony are the best adapted for bearings. 

Mr. Clamer, however, does not agree with this, and says that alloys 
containing below 13 per cent antimony can likewise be said to consist 
of hard grains consisting of the composition itself, imbedded in the 
softer material, lead, as mentioned above. He says: “It has been my 
experience that, although the friction may be higher in such alloys, 
the wear is greatly diminished, and where pressures are light, causing 
no deformation, this is a great advantage. I have seen many instances 
in service where alloys between 15 per cent and 25 per cent w T ere 
greatly inferior to alloys between 8 per cent and 12 per cent, owing to 
their frequent renewal due to w r ear. It will perhaps be interesting to 
hear that the Pennsylvania Railroad Company, at the suggestion of 
Dr. Dudley, their chemist, have adopted the 13 per cent antimonial 
lead alloy as a filling metal for bearings in order to obtain the best 
results. In a general way my own work in the subject has confirmed 
the opinion that lead is the best wear-resisting metal known, and 
that with increasing antimony, or increasing hardness and brittleness, 
the wear becomes more marked. This is due to the splitting up of 
the harder particles.” 

The friction, as we may naturally expect, becomes less w T ith increase 
of antimony, and the temperature of running likewise diminished when 
running under normal conditions; but the harder the alloy, the more 
difficulty is experienced in bringing it primarily to a perfect bearing, 
and the greater the liability of heating through aggravated conditions. 
The wear on the journal one would naturally expect to be decreased 
with increasing hardness; but this journal wear is in all probability 
not due so much to the alloy directly as it is to the fact that the softer 
metals collect grit, principally from the small particles of steel from 
the worn journal, and, acting as a lap, cause rapid wear. With the 
harder metals these particles are worked out without becoming im¬ 
bedded. 

The cost of the lead and antimony alloy is the least which can be 
produced. It can be used in many services where higher priced alloys 
are being relied upon mainly for their high cost. It is one of the 
greatest extravagances of large industrial establishments to use ma¬ 
terials that are too good for certain uses, and even perhaps unsuited, 
under the supposition that they must be good because they paid a good 
price for them. This fact has no greater exemplification than in the 
purchase of babbitt metal, and is due to the great uncertainty which 
exists, not only among consumers, but among manufacturers, many of 
whom carry on their business much the same as the patent-medicine 
man. 

2. Lead, Antimony, and Tin: It should not be assumed that anti¬ 
mony-lead is the cheapest alloy to use under all circumstances; not so, 
for when high pressures are to be encountered, tin is a very desirable 


43 


ALLOYS FOR BEARINGS 

adjunct. Tin imparts to the lead-antimony alloy rigidity and hard¬ 
ness without increasing brittleness, and can produce alloys of sufficient 
compressive strength for nearly all uses. The structure of a triple 
alloy of this nature is quite complicated, and not yet sufficiently defined. 

The cost of the alloy increases with increase of tin; but for certain 
uses, where sufficient compressive strength cannot be gotten by anti¬ 
mony, because of its accompanying brittleness, it is indispensable, and 
will answer in nearly every case where the tin basis babbitts are used. 

3. Tin and Antimony: These are seldom used alone as bearing 
alloys, but are extensively used for so-called Brittania ware, and in 
equal proportions for valve seats, etc. 

4. Tin, Antimony, and Copper: This combination is what is known 
as genuine babbitt, after its inventor, Isaac Babbitt, who presumably 
was the first man to conceive the idea of lining bearings with fusible 
metal. The formula, which for no arbitrary reason he recommended, 


is as follows: 

Tin. 89.1 

Antimony . 7.4 

Copper . 3.7 


This formula is still considered the standard of excellence in the 
trade, and has been adopted by many of the leading railroads, the 
United States Government, and many industrial establishments. It is 
used in the majority of cases w r here cheaper composition would do 
equally as well. It is the most costly of all bearing alloys because of 
the high content of tin. 

5. Tin-Antimony-Lead-Copper: This quadruple combination of metals 
cannot be satisfactorily described, as it would no doubt take years 
of study to fathom the complicity of the metallic combinations here 
represented. Suffice it to say that lead, although of itself a soft metal, 
renders this alloy, when added in but small proportions, harder, stiffer, 
more easily melted and superior in every way to the alloy without it, 
and yet consumers 'will raise their hands in horror when a trifling 
percentage of lead is found in their genuine babbitt. This is one 
of the instances where cheapening of the product is beneficial. 

The foregoing represents the more important combinations of alloys 
of tin and lead basis. These are of far more importance in the arts 
than the white metals, the main portion or basis of which is zinc. 

At various times new combinations of zinc have been proposed, but, 
with very few exceptions, they have not come into popular use for 
two reasons: First, because of the great tendency of zinc to adhere to 
iron when even slightly heated. What is technically known as gal¬ 
vanizing the journal is effected under these conditions. Second, be¬ 
cause of the brittleness produced under the effects of heat, such as is 
produced by friction when lubrication is -interfered with, and conse¬ 
quent danger of breakage. 

Bronzes 

Bronze is the term which originally was applied to alloys of copper 
and tin as distinguished from alloys of copper and zinc. 





44 


No. ii—BEARINGS 


1. Copper and Tin: This, according to our general conception of the 
word, is a bronze only when the copper content exceeds that of the 
tin. According to the proportions in which the metals exist, it has 
widely different properties. In general, the alloy hardens when tin is 
present up to proportions of 30 per cent or a little over, and when this 
limit is exceeded, it takes on more and more the nature of tin until 
pure tin is reached. From a scientific point of view this alloy is 
one of the most interesting, and has attracted the attention of many 
investigators, who have spent years of study on it, to learn its various 
properties and explain its constitution. 

The alloys which interest us most, however, are those which are so 
constituted as to be adapted for bearing purposes. These would be 
said to contain from 3 to 15 per cent tin, and from 85 to 97 per cent 
copper. The alloy of tin containing a small percentage of copper is 
often used as a babbitt metal, but this comes under the class of white 
metals, which have already been discussed. Bronze containing above 



COPPER 

AND TIN, AND 

COPPER, 

TIN, AND 

LEAD SERIES 


Copper 

Tin 

Lead 

Friction 

Temp. 

above 

Room 

Wear in 
Grams 

1 

85.76 

14.90 


13 

50 

.2800 

2 

90.67 

9.45 


13 

51 

.1768 

3 

95.01 

4.95 


16 

52 

.0776 

4 

90.82 

4.62 

4.82 

14 

53 

.0542 

5 

85.12 

4.64 

10.64 

is y 2 

56 

.0380 

6 

81.27 

5.17 

14.14 

18% 

58 

.0327 

7 

75? 

5? 

20? 

18y 2 

58 

.0277 

8 

68.71 

5.24 

26.67 

18 

58 

.0204 

9 

64.34 

4.70 

31.22 

18 

44 

.0130 


15 per cent of tin has been recommended at various times for bearings, 
owing to its hardness, but very unwisely, for such a bearing demands 
mechanical perfection and perfect lubrication. It has no plasticity of 
its own, and as soon as the oil film is interrupted, rapid abrasion and 
“gripment” take place, with hot boxes as the result. The very erro¬ 
neous idea is still held by many, that to resist wear and run with the 
least possible friction, a bearing alloy must be as hard as possible. It 
is true that hard bodies in contact move with less friction than soft 
ones; but the alloy which is the least liable to heat and cause trouble is 
the one which will stand the greatest amount of ill use; by this is 
meant an alloy which has sufficient plasticity to adapt itself to the 
irregularities of service without undue wear. 

The alloys of copper and tin were used extensively some twenty or 
twenty-five years ago, and were considered the standard for railroad 
and machinery bearings. The old alloy, known as “Cannon Bronze,” 
containing 7 parts copper and 1 part tin, is still being specified by some 
few unprogressive railroad men and machinery builders. 

2. Copper , Tin, and Lead: This composition is now the recognized 
standard bearing bronze, its advantage over the bi-compound coming 
from the introduction of lead. The bronze containing lead is less liable 
to heat under the same state of lubrication, etc., and the rate of wear 
is much diminished. For these reasons and the additional fact that 






ROLLER BEARINGS 


45 


lead is cheaper than tin, it seems desirable to produce a hearing metal 
with as much lead and as little tin as possible. The metal known as 
Ex. E. composition (tin 7 per cent, lead 15 per cent, copper 78 per cent) 
is stated to be the best that can be devised. This alloy contains the 
smallest quantity of^in that w T ill hold the lead alloyed with the 
copper. By adding a small percentage of nickel, however, to the extent 
of one-half to 1 per cent, a larger proportion of lead may be used, and 
successful bronzes have been made by this process, which contained 
as much as 30 per cent lead. Such bronzes, containing a large amount 
of lead, through the addition of nickel, are known in the trade as 
“Plastic Bronzes” and are a regular commercial article. The table on 
page 44 gives the results of tests on different compositions of bronzes. 


CHAPTER VII 


FRICTION OF ROLLER BEARINGS* 

During the years 1904-05 a series of tests on roller bearings was 
conducted at the Case School of Applied Science, Cleveland. A complete 
report of these tests was published by Professor C. H. Benjamin in 
the October, 1905, issue of Machinery, of which the following is an 
abstract. An attempt was made in these experiments to compare roller 
bearings with plain cast iron bearings and with babbitted bearings 
under similar conditions. Four sizes of bearings were used in the 
tests, measuring respectively 1 15-16, 2 3-16, 2 7-16 and 215-16 inches 
in diameter. The lengths of journals were four times the diameters. 

The bearings were in two parts and were held in a circular yoke by 
setscrews. This yoke carried two vertical spindles, one above and one 
below, on which were placed the weights for loading the bearings. 
The friction was measured either by the deflection of the compound 
pendulum thus formed, or, as in most of the experiments, by weigh¬ 
ing its tendency to deflect by means of an attached cord running over 
a pulley and carrying a scale pan, as shown in Fig. 19. The shafts 
or journals used were of ordinary machinery steel, carefully turned to 
size and having a smooth finish. These shafts were rotated at the 
speeds shown by means of a belt and pulley. The cast-iron bearings 
used for comparison were cast whole and bored to size, but the bab¬ 
bitted ones were in halves and were held the same as the roller bear¬ 
ings. 

In beginning an experiment, a pointer on the lower end of the pen¬ 
dulum was brought to a zero mark vertically beneath the center of the 
shaft by adjusting the screws in the yoke. After the shaft began to 


* Machinery, October, 1905. 



46 


No. ii—BEARINGS 


revolve the pointer was held to the zero mark by putting weights on 
the scale pan. The product of the force thus applied to the pendulum 
by the distance of the point of application from the center of shaft 
gave the moment of friction, and dividing this by the radius of jour¬ 
nal gave the friction at the surface of the journal. Dividing this again 
by the total weight on the journal gave the coefficient of friction. 



Fig'. 19. Apparatus for Testing Roller Bearings 


In the first set of experiments Hyatt roller bearings were compared 
with plain cast iron sleeves at a uniform speed of 480 revolutions per 
minute, and under loads varying from 64 to 264 pounds. The cast 
iron bearings were thoroughly and copiously oiled, the lubrication be¬ 
ing rather better than would be the case in ordinary practice. Table 
I shows the results of the test on one bearing in detail, and from this 
it is seen that the value of /, the coefficient of friction, diminishes as 





















































47 


ROLLER BEARINGS 


the load increases, or in other words, the friction did not increase 
as fast as the load. This holds true as a general rule in all the roller 
bearings, but not generally in the plain bearings, either cast iron or 
babbitt. 

Table II gives a summary of this series of experiments for the 
different sizes of journals, the different loads being the same as in 
Table I. The relatively high values of / in the 2 3/16 and 2 15/16 roller 


TABLE I. 

Journal 115*13 Inches In Diameter. 480 Revolutions per minute. 


Total Load, 
pounds. 

Friction. 

Values of /. 

Hyatt 

Plain. 

Hyatt. 

Plain. 

64.2 

2.84 

10.24 

.036 

.160 

114.2 

3.27 

12.10 

.029 

.106 

164.2 

4.21 

19.10 

.026 

.116 

214.2 

4.78 

22.35 

.022 

.104 

264.2 

5.15 

26.10 

.019 

.099 

Average 



.026 

.117 


bearings were due to the snugness of the fit between the journal and 
the bearing, and show the advisability of as easy a fit as in ordinary 
bearings. 

The same Hyatt bearings were used in the second set of experiments, 
but were compared with the McKeel solid roller bearings and with 
plain babbitted bearings freely oiled. The McKeel bearings contained 
rolls turned from solid steel and guided by spherical ends fitting re- 

TABLE II. 

Values of Coefficient of Friction f. Speed 480 Revolutions per minute. 


Diameter 

of 

Journal. 

Hvatt Bearing. 

Plain Bearing. 

Max. 

Min. 

Ave. 

Max. 

Min. 

Ave. 

HI 

2* 

.036 

.052 

.041 

.053 

.019 

.084 

.025 

.049 

.026 

.040 

.030 

.051 

.160 

.129 

.143 

.138 

.099 

.071 

.076 

.091 

.117 

.094 

.104 

.104 


cesses in cage rings at each end. The cage rings were joined to each 
other by steel rods parallel to the rolls. The same apparatus was 
used as in the former tests, but heavier loads were used and the ma¬ 
chine was run at a slightly higher speed. Table III shows the detailed 
results of experiments on one size of journal, and is similar to Table I. 
The last value given for the Hyatt bearing shows distortion of the 
roller due to the load and indicates the limit for this size. This is 
omitted in getting the averages. There is the same indication as in 
Table I of a decrease of f with increase of load, and this was noticed 
in all the tests. The results for the babbitt metal are not as uni¬ 
form as the others on account of the difficulty of balancing. 























No. n—BEARINGS 


MOV 23 1912 

48 


Under a load of 358.3 pounds the solid roller bearing showed an 
end thrust of about 20 pounds, which would account for the difference 
in friction between that and the Hyatt. Table IV gives a summary 
of the testa in this series and may be compared with Table II. The 
relatively high values for the Hyatt 2 7-16 bearing must be due to a 
slight cramping of the rolls due to too close a fit, as was noted in some 
of the former experiments. Under a load of 470 pounds, the Hyatt 

table in. 

Journal 115-16 Inches in Diameter. Speed of 500 Revolutions per minute. 


Total Load. 

Friction. 

Value of/. 

Hyatt. 

.McKeel. 

Babbitt. 

Hyatt. 

McKeel. 

Babbitt. 

113.3 

3.64 

8.77 

8.38 

.032 

.033 

.074 

162.3 

3.77 

4.24 

8.97 

.023 

.026 

.055 

211.3 

4.04 

5.24 

8.97 

.019 

.025 

.042 

260.3 

4.3t 

5.37 

8.97 

.016 

.021 

.034 

309.3 

4.57 

6.46 

10.13 

.015 

.021 

.033 

358.3 

4.71 

6.73 

10.75 

.013 

.019 

,030 

407.3 

4.84 

7.27 

11.98 

.012 

.018 

.029 

456.3 

37,70 

7.81 

20.90 


.017 

.046 

Averages 




.0186 

.0225 

.043 


bearings developed an end thrust of 13.5 pounds and the McKeel one 
of 11 pounds. This end thrust is due to a slight skewing of the rolls 
and would vary, sometimes even reversing in direction. 

The babbitt bearing is a slight improvement over the cast-iron sleeve, 
but the difference is quite as apt to be due to improved lubrication. 
(Notice the variation in the averages for the various sizes in Table IV). 

In conclusion it may be said that the friction of the roller bearing 

TABLE IV. 

Values of Coefficient of Friction f. Speed 660 Revolutions per minute. 


Diameter 
of Journal. 

Hyatt Bearing. 

McKeel Bearing. 

Babbitt Bearing. 

Max. 

Min. 

Ave. 

Mar. 

Min 

Ave. 

Max. 

Min. 

Ave. 

HI 

.032 

.012 

.018 

.033 

.017 

.022 

.074 

.029 

.043 

s* 

.019 

.011 

.014 




.088 

.078 

.082 

.042 

.025 

.032 

.028 

.015 

.021 

.114 

.083 

.096 

015 

.029 

.022 

.025 

.039 

.019 

.027 

.125 

.089 

.107 


is shown to be from one-fifth to one-third that of a plain bearing at 
moderate loads and speeds. It is also noticeable that as the load on a 
roller bearing increases the coefficient of friction decreases. It was 
found by the experimenters that a slight change in the pressure due 
to the adjusting nuts was sufficient to increase the friction consider¬ 
ably. In the McKeel bearing the rolls bore on a cast-iron sleeve and 
in the Hyatt on a soft steel one. If roller bearings are properly ad¬ 
justed and not overloaded, a saving of from 2/3 to 3/4 of the friction 
may be reasonably expected. 































Iffo. 50. Principles and Practice oiA •- 
sembling Machine Tools, Part I. 

No. 51. Principles and Practice of As¬ 
sembling' Machine Tools, Part II. 

No. 52. Advanced Shop Arithmetic for 
the Machinist. 

No. 53. Use of Logarithms and Logar¬ 
ithmic Tables. 

No. 54. Solution of Triangles, Part I. 
—Methods, Rules and Examples. 

No. 55. Solution of Triangles, Part II. 
—Tables of Natural Functions. 

No. 56. Ball Bearings.— Principles of 
Design and Construction. 

No. 57. Metal Spinning.— Machines, 
Tools and Methods Used. 

No. 58. Helical and Elliptic Springs.— 
Calculation and Design. 

No. 59. Machines, Tools and Methods 
of Automobile Manufacture. 

No. 60. Construction and Manufacture 
of Automobiles. 

No. 61. Blacksmith Shop Practice.— 

Model Blacksmith Shop; Welding; Forg¬ 
ing of Hooks and Chains; Miscellaneous. 

No. 62. Hardness and Durability Test¬ 
ing of Metals. 

No. 63. Heat Treatment of Steel.— 

Hardening, Tempering, Case-Hardening. 


No. 

64. 

Gage Making and 

Lapping. 

No. 

65. 

Formulas and 

Constants for 

Gas Engine Design. 



No. 

66. 

Heating and 

Ventilation of 

Shops 

and 

Offices. 



No. 

67. 

Boilers. 



No. 

68. 

Boiler Furnaces 

and Cliim- 


neys. 

No. 69. Peed Water Appliances. 


No. 70. Steam Engines. 

No. 71. Steam Turbines. 

No. 72. Pumps, Condensers, Steam and 
Water Piping. 

No. 73. Principles and Applications of 
Electricity, Part I.—Static Electricity; 
Electrical Measurements; Batteries. 

No. 74. Principles and Applications of 
Electricity, Part II. — Magnetism; Elec¬ 
tro-Magnetism; Electro-Plating. 

No. 75. Principles and Applications of 
Electricity, Part III. — Dynamos; Motors; 
Electric Railways. 

No. 76. Principles and Applications of 
Electricity, Part IV. — Electric Lighting. 

No. 77. Principles and Applications of 
Electricity, Part V. — Telegraph and Tele¬ 
phone. 

No. 78. Principles and Applications of 
Electricity, Part VI.—Transmission of 
Power. 


No. 79. Locomotive Building, Part I.— 

Main and Side Rods. 

No. 80. Locomotive Building, Part II. 
—Wheels; Axles; Driving Boxes. 

No. 81. Locomotive Building, Part III. 
■—Cylinders and Frames. 

No. 82. Locomotive Building, Part IV. 
—Valve Motion. 

No. 83. Locomotive Building, Part V. 

—Boiler Shop Practice. 

No. 84. Locomotive Building, Part VI. 
—Erecting. 

No. 85. Mechanical Drawing, Part I. 

—Instruments; Materials; Geometrical 
Problems. 

No. 86. Mechanical Drawing, Part II. 

—Projection. 

No. 87. Mechanical Drawing, Part III. 

—Machine Details. 

No. 88. Mechanical Drawing, Part IV. 

—Machine Details. 

No. 89. The Theory of Shrinkage and 
Forced Fits. 

No. 90. Railway Repair Shop Practice. 

No. 91. Operation of Machine Tools.— 

The Lathe, Part I. 

No. 92. Operation of Machine Tools.— 

The Lathe, Part II. 

No. 93. Operation of Machine Tools.— 

Planer, Shaper, Slotter. 

No. 94. Operation of Machine Tools.— 

Drilling Machines. 

No. 95. Operation of Machine Tools.— 

Boring Machines. 

No. 96. Operation of Machine Tools.— 
Milling Machines, Part I. 

No. 97. Operation of Machine Tools.— 
Milling Machines, Part II. 

No. 98. Operation of Machine Tools.— 
Grinding Machines. 

No. 99. Automatic Screw Machine 
Practice, Part I. — Operation of the Brown 
& Sharpe Automatic Screw Machine. 

No. 100. Automatic Screw Machine 

Practice, Part II. — Designing and Cutting 
Cams for the Automatic Screw Machine. 
. No. 101. Automatic Screw Machine 

Practice, Part III.—Circular Forming and 
Cut-off Tools. 

No. 102. Automatic Screw Machine 

Practice, Part IV. — External Cutting 
Tools. 

No. 103. Automatic Screw Machine 

Practice, Part V — Internal Cutting Tools. 

No. 104. Automatic Screw Machine 

Practice, Part VI. — Threading Operations. 

No. 105. Automatic Screw Machine 

Practice. Part VII. — Knurling Operations. 

No. 106. Automatic Screw Machine 

Practice, Part VIII.—Cross Drilling, Burr¬ 
ing and Slotting Operations. 


ADDITIONAL TITLES WILL BE ANNOUNCED IN MACHINERY FROM TIME TO TIME 


MACHINERY’S DATA SHEET SERIES 

Machinery’s Data Sheet Books include the well-known series of Data Sheets 
originated by Machinery, and issued monthly as supplements to the publication; 
of these Data Sheets over 500 have been published, and 6,000,000 copies sold. Re¬ 
vised and greatly amplified, they are now presented in book form, kindred sub¬ 
jects being grouped together. The purchaser may secure either the books on 
those subjects in which he is specially interested, or, if he pleases, the whole set at 
one time. The price of each book is 25 cents (one shilling) delivered anywhero 
in the world. 




























f 



0 033 266 533 . 


CONTENTS OF DATA SHEET BOOKS 


No. 1. Screw Threads. —United States, 
Whitworth, Sharp Y- and British Associa¬ 
tion Standard Threads; Briggs Pipe 
Thread; Oil Well Casing Gages; Fire Hose 
Copnections; Acme Thread; Worm 
Threads; Metric Threads; Machine, Wood, 
and Lag Screw Threads; Carriage Bolt 
Tlireads, etc. 

No. 2. Screws, Bolts and Nuts. —Fil¬ 
lister-head, Square-head, Pleadless, Col¬ 
lar-head and Hexagon-head Screws; Stand¬ 
ard and Special Nuts; T-nuts, T-bolts and 
Washers; Thumb Screws and Nuts; A. L. 
A. M. Standard Screws and Nuts; Machine 
Screw Heads; Wood Screws; Tap Drills; 
Lock Nuts; Eye-bolts, etc. 

No. 3. Taps and Dies. —Hand, Machine, 
Tapper and Machine Screw Taps; Taper 
Die Taps; Sellers Plobs; Screw Machine 
Taps; Straight and Taper BoiDr Taps; 
Stay-bolt, Washout, and Patch-bolt Taps; 
Pipe Taps and Hobs; Solid Square, Round 
Adjustable and Spring Screw Threading 
Dies. 

No. 4. Reamers, Sockets, Drills and 
Milling Cutters. —Hand Reamers; Shell 
Reamers and Arbors; Pipe Reamers; Taper 
Pins and Reamers; Brown & Sharpe, 
Morse and Jarno Taper Sockets and Ream¬ 
ers; Drills; Wire Gages; Milling Cutters; 
Setting Angles for Milling Teeth in End 
Mills and Angular Cutters, etc. 

No. 5. Spur Gearing. —Diametral and 
Circular Pitch; Dimensions of Spur Gears; 
Tables of Pitch Diameters; Odontograph 
Tables; Rolling Mill Gearing; Strength of 
Spur Gears; Horsepower Transmitted by 
Cast-iron and Rawhide Pinions; Design of 
Spur Gears; Weight of Cast-iron Gears; 
Epicyclic Gearing. 

No. 6. Bevel, Spiral and Worm Gear¬ 
ing. —Rules and Formulas for Bevel 
Gears; Strength of Bevel Gears; Design 
of Bevel Gears; Rules and Formulas for 
Spiral Gearing; Tables Facilitating Calcu¬ 
lations; Diagram for Cutters for Spiral 
Gears; Rules and Formulas for Worm 
Gearing, etc. 

No. 7. Shafting, Keys and Keyways.— 

Horsepower of Shafting; Diagrams and 
Tables for the Strength of Shafting; 
Forcing, Driving, Shrinking and Running 
Fits; Woodruff Keys; United States Navy 
Standard Keys; Gib Keys; Milling Key- 
ways; Duplex Keys. 

No. 8. Bearings, Couplings, Clutches, 
Crane Chain and Hooks. —Pillow Blocks; 
Babbitted Bearings; Ball and Roller Bear¬ 
ings; Clamp Couplings; Plate Couplings; 
Flange Couplings; Tooth Clutches; Crab 
Couplings; Cone Clutches; Universal 
Joints; Crane Chain; Chain Friction; 
Crane Hooks; Drum Scores. 

No. 9. Springs, Slides and Machine 
Details. —Formulas and Tables for Spring 
Calculations; Machine Slides; Machine 
Handles and Levers; Collars; Hand 
Wheels; Pins and Cotters; Turn-buckles, 
etc. 

No. 10. Motor Drive, Speeds and Peeds, 
Chang'e Gearing, and Boring Bars. —Power 
required for Machine Tools; Cutting 
Speeds and Feeds for Carbon and High¬ 
speed Steel; Screw Machine Speeds and 
Feeds; Heat Treatment of High-speed 


Stbel Tools; Taper Turning; Change Gear¬ 
ing for the Lathe; Boring Bars and Tools, 
etc. 

No. 11. Milling Machine Indexing, 
Clamping Devices and Planer Jacks.— 

Tables for Milling Machine Indexing; 
Change Gears for Milling Spirals; Angles 
for setting Indexing Head when Milling 
Clutches; Jig Clamping Devices; Straps 
p.nd Clamps; Planer Jacks. 

No. 12. Pipe and Pipe Fittings. —Pipe 
Threads and Gages; Cast-iron Fittings; 
Bronze Fittings; Pipe Flanges; Pipe 
Bends; Pipe Clamps and Hangers; Dimen¬ 
sions of Pipe for Various Services, etc. 

No. 13. Boilers and Chimneys. —Flue 
Spacing and Bracing for Boilers; Strength 
of Boiler Joints; Riveting; Boiler Setting; 
Chimneys. 

No. 14. Locomotive and Railway Data. 

—Locomotive Boilers; Bearing Pressures 
for Locomotive Journals; Locomotive 
Classifications; Rail Sections; Frogs, 
Switches and Cross-overs; Tires; Tractive 
Force; Inertia of Trains; Brake Levers; 
Brake Rods, etc. 

No. 15. Steam and Gas Engines. —Sat¬ 
urated Steam; Steam Pipe Sizes; Steam 
Engine Design; Volume of Cylinders; 
Stuffiing Boxes; Setting Corliss Engine 
Valve Gears; Condenser and Air Pump 
Data; Horsepower of Gasoline Engines; 
Automobile Engine Crankshafts, etc. 

No. 16. Mathematical Tables. —Squares 
of Mixed Numbers; Functions of Frac¬ 
tions; Circumference and Diameters of 
Circles; Tables for Spacing off Circles: 
Solution of Triangles; Formulas for Solv¬ 
ing Regular Polygons; Geometrical Pro¬ 
gression, etc. 

No. 17. Mechanics and Strength of Ma¬ 
terials. —Work; Energy; Centrifugal 
Force; Center of Gravity; Motion; Fric¬ 
tion; Pendulum; Falling Bodies; Strength 
of Materials; Strength of Flat Plates; 
Ratio of Outside and Inside Radii of 
Thick Cylinders, etc. 

No. 18. Beam Formulas and Structural 
Design. —-Beam Formulas; Sectional Mod¬ 
uli of Structural Shapes; Beam Charts; 
Net Areas of Structural Angles; Rivet 
Spacing; Splices for Channels and I- 
beams; Stresses in Roof Trusses, etc. 

No. 19. Belt, Rope and Chain Drives.— 
Dimensions of Pulleys; Weights of Pul¬ 
leys; Horsepower of Belting; Belt Veloc¬ 
ity; Angular Belt Drives; Horsepower 
transmitted by Ropes; Sheaves for Rope 
Drive; Bending Stresses in Wire Ropes; 
Sprockets for Link Chains; Formulas and 
Tables for Various Classes of Driving 
Chain. 

No. 20. Wiring Diagrams, Heating and 
Ventilation, and Miscellaneous Tables.— 

Typical Motor Wiring Diagrams; Resist¬ 
ance of Round Copper Wire; Rubber Cov¬ 
ered Cables; Current Densities for Vari¬ 
ous Contacts and Materials; Centrifugal 
Fan and Blower Capacities; Hot Water 
Main Capacities; Miscellaneous Tables: 
Decimal Equivalents, Metric Conversion 
Tables, Weights and Specific Gravity of 
Metals, Weights of Fillets, Drafting-room 
Conventions, etc. 


Machinery, the monthly mechanical journal, originator of the Reference and 
Data Sheet Series, is published in three editions—the Shop Edition, $1.00 a year; 
the Engineering Edition, $2.00 a year, and the Foreign Edition, $3.00 a year. 

The Industrial Press, Publishers of Machinery, 

49-55 Lafayette Street, New York City, U. S. A, 










